WORKS   OF 
PROFESSOR  A.  M.  GREENE, 

PUBLISHED    BY 

JOHN  WILEY  &  SONS,  Inc. 


Elements  of  Heating  and  Ventilation. 

A  Text-book  for  Technical  Students  and  a 
Reference  Book  for  Engineers.  8vo,  vi  +324 
pages,  223  figures.  Cloth,  $2.50  net. 

Pumping  Machinery. 

A  Treatise  on  the  History,  Design,  Construction, 
and  Operation  of  Various  Forms  of  Pumps. 
8vo,  vi  +703  pages,  504  figures.  Cloth,  $4.00  net. 

The  Elements  of  Refrigeration. 

A  Text-book  for  Students,  Engineers  and  Ware- 
housemen, vi  +472  pages.  6  by  9.  192  figures. 
Cloth,  $4.00  net. 

BY  SPANGLER,  GREENE,  AND  MARSHALL: 

Elements  of  Steam  Engineering. 

Third  Edition,  Revised.  8vo,  v+296  pages, 
284  figures.  Cloth,  $3.00. 


PUMPING  MACHINERY 

A  TREATISE   ON   THE 

HISTORY,    DESIGN,    CONSTRUCTION 

AND   OPERATION    OF  VARIOUS 

FORMS   OF    PUMPS 


BY 

ARTHUR    M.    GREENE,    JR. 

Professor  of  Mechanical  Engineering,  Russell  Sage  Foundation,  Rensselaer  Polytechnic 
Institute;  Sometime  Junior  Dean,  School  of  Engineering ,  University  of  Missouri 


SECOND    EDITION,   REVISED 


NEW  YORK 

JOHN  WILEY  &  SONS,  INC. 

LONDON:    CHAPMAN  &  HALL,  LIMITED 

1919 


Engineering 
Library 


Copyright,  1911,  1918,  1919, 

BY 

ARTHUR   M.    GREENE,   JR. 


PRESS    OF 

BRAUNWORTH    &    CO. 

BOOK    MANUFACTURER* 

BROOKLYN.     N.    Y. 


PREFACE  TO  SECOND  EDITION 


IN  the  second  edition  of  Pumping  Machinery  it  has  been 
thought  advisable  to  rearrange  the  treatment  of  water  valve 
design  and  the  dynamics  of  the  water  end.  The  air  lift  pump 
has  been  analyzed  in  a  different  manner  to  agree  more  closely 
with  the  true  variation  of  pressure.  Other  small  minor  changes 
have  been  made. 

A.  M.  G.,  Jr. 

Sunnyslope,  Troy,  N.  Y. 
May,  1918. 


lii 


PREFACE 


SEVEN  YEARS  ago  the  course  of  lectures  which  forms  the 
basis  of  this  book  was  given  as  a  required  course  to  the  mechan- 
ical engineering  students  at  the  University  of  Missouri.  As  a 
number  of  civil  engineering  students  elected  this,  it  occurred 
to  the  author  that  there  was  a  demand  for  a  book  to  cover 
the  field  as  he  had  planned  the  course. 

The  general  plan  of  the  work  is  to  give  a  brief  historical 
outline  of  the  development  of  pumping  machinery  which  is 
closely  connected  with  the  history  of  the  steam  engine;  then 
to  examine  and  describe  the  action  of  a  number  of  common 
forms  of  pumps,  following  this  by  the  methods  of  design  of 
pumping  apparatus. 

The  book  is  intended  for  the  use  of  those  who  have  studied 
hydraulics,  mechanics,  and  the  strength  of  materials.  It  is 
intended  to  develop  certain  general  principles  of  mechanics 
which  are  applicable  to  pumping  machinery  as  well  as  to  train 
the  student  in  application  of  certain  of  the  theoretical  portions 
of  the  studies  of  an  engineering  course. 

The  first  two  chapters  are  devoted  to  the  development  of 
pumping  machinery,  because  the  author  feels  the  need,  of  all 
engineers,  of  a  more  intimate  knowledge  of  the  great  works  of 
the  early  members  of  the  profession.  The  pump  is  so  closely 
allied  with  the  steam  engine  in  early  times  that  this  history 
traces  the  early  forms  of  that  machine.  The  historical  works 
mentioned  in  the  list  of  text-books  in  the  bibliography  as  well  as 
the  historical  articles  given  in  the  same  catalogue  have  been 
indispensable  in  the  preparation  of  these  chapters. 

The  descriptive  chapters  on  modern  pumps  have  been 
prepared  from  the  catalogues  and  bulletins  of  manufacturers 
and  from  current  technical  literature.  The  figures  have  been 
redrawn  from  these  sources  to  fit  them  for  this  work.  The 
half-tones  have  been  made  from  photographs  or  cuts  of  publi- 
cations kindly  furnished  by  the  various  manufacturers.  The 


VI 


PREFACE 


thanks  of  the  author  to  these  manufacturers  are  here  publicly 
expressed.  The  various  articles  in  the  technical  press  which 
have  been  of  assistance  in  the  preparation  of  the  lectures  and 
this  work  are  listed  as  bibliography.  The  author  is  indebted 
to  the  writers  of  these  articles  and  texts  and  to  their  publishers 
for  the  aid  he  has  received.  These  chapters  have  been  incor- 
porated into  the  book  with  certain  tables  from  the  catalogues, 
so  that  the  student  may  have  in  his  library  a  reference  book 
which  will  not  only  give  him  the  peculiarities  of  the  different 
forms  of  special  pumps,  but  the  dimensions  or  other  data  may 
be  found  so  that  he  may  know  what  size  of  pump  may  be  had 
for  a  certain  purpose. 

In  the  preparation  of  the  theoretical  part  of  the  work  of 
design  the  author  has  used  the  methods  employed  by  him  in 
courses  in  steam-engine  design  and  hydraulic  motors,  adding 
thereto  methods  developed  from  "Die  Pumpen,"  by  Hartmann- 
Knoke  and  Berg,  and  "  Die  Zentrifugal  Pumpen,"  by  Fritz 
Neumann.  These  two  books  have  been  of  great  assistance  in 
the  preparation  of  the  work.  The  theory  relating  to  air-lift 
pumps  differs  from  that  given  by  Harris  in  his  articles  in  the 
Proceedings  of  the  American  Society  of  Civil  Engineers,  which 
were  originally  considered. 

The  author  especially  wishes  to  express  his  thanks  in  this 
place,  to  his  wife,  Mary  E.  Lewis  Greene,  for  her  aid  in  the 
preparation  of  the  manuscript  and  in  proof-reading,  as  well  as 
for  her  advice  during  the  progress  of  the  work. 

Grateful  acknowledgments  are  made  to  the  following: 
International  Steam  Pump  Company  and  to  Mr.  William 
Schwanhauser,  Henry  R.  Worthington  Company,  the  Allis- 
Chalmers  Company,  and  Mr.  Win  Sando,  the  Snow  Pump  Works, 
The  Holly  Pump  Works,  and  Mr.  Decrow,  The  American  Steam 
Pump  Company,  the  Alberger  Company,  the  Buffalo  Pump 
Company,  the  Prescott  Pump  Company,  Deane  Brothers  Pump 
Company,  the  National  Board  of  Fire  Underwriters,  Mr.  Francis 
Head,  Prof.  Lewis  F.  Moody,  Mr.  F.  C.  Dunlap,  Mr.  Lester 
French,  and  Mr.  A.  F.  Rolf  for  their  aid  in  furnishing  data., 
illustrations  and  other  materials. 

THE  AUTHOR 

TROY,  March  15,  1911. 


TABLE   OF    CONTENTS 


CHAPTER  I 

PAGB 

HISTORICAL  DEVELOPMENT i 

CHAPTER  II 
RECENT  HISTORY 56 

CHAPTER  III 
MODERN  RECIPROCATING  PUMPS : 130 

CHAPTER  IV 
SIMPLEX  PUMPS 166 

CHAPTER  V 
DYNAMICS  OF  WATER  END , 190 

CHAPTER  VI 
DESIGN  OF  PARTS 260 

CHAPTER  VII. 
DYNAMICS  OF  STEAM  END 324 

CHAPTER  VIII 
STEAM  END  DETAILS  .  • 344 

CHAPTER  IX 
TEST  OF  PUMPING  ENGINES T 410 

CHAPTER  X 

HIGH  DUTY  PUMPS  AND  WATER  WORKS  STATIONS 428 

vii 


viii  TABLE  OF   CONTENTS 

CHAPTER  XI 

PAGE 

SPECIAL  PUMPING  MACHINERY 456 

CHAPTER  XII 
INJECTOR  AND  PULSOMETER 503 

CHAPTER  XIII 
AIR-LIFT  PUMPS  AND  PNEUMATIC  PUMPS 512 

CHAPTER   XIV 
CENTRIFUGAL  PUMPS 535 

CHAPTER  XV 
MINE  PUMPS 659 


PUMPING  MACHINERY 


CHAPTER  I 

HISTORICAL   DEVELOPMENT 

THE  pump,  considered  as  any  apparatus  used  to  raise  water, 
is  one  of  the  oldest  machines  known  to  man.  Long  before 
the  Christian  Era,  when  man  emerged  from  the  age  of  the  hunter 
into  that  of  the  shepherd,  he  found  it  necessary  to  raise  water  from 
wells  for  his  flocks  in  places  where  there  were  no  pure  streams. 
Indeed,  even  before  this  age  there  must  have  been  devices 
to  raise  water  from  low  levels  tc  supply  the  personal  needs 
of  the  hunter  and  those  dependent  on  him. 

Following  the  age  of  the  shepherd  came  that  of  the  farmer, 
when  the  demand  for  an  apparatus  to  raise  water  was  greater 
than  before.  The  shepherd  no  longer  wandered  over  the 
country  in  search  of  pasture,  but  now  he  cared  for  a  definite 
tract  of  land  to  furnish  the  food  supply  for  his  flocks,  his  herds, 
and  his  family.  Unfortunately  it  became  necessary  for  him 
to  use  lands  on  which  there  was  not  a  sufficient  natural  supply 
of  water  for  irrigation,  and  he  was  compelled  to  lift  water  from 
low  streams  to  these  fields,  so  as  to  increase  the  yield  from  his 
land. 

The  next  age  was  that  of  townsman  and  manufacturer. 
For  protection,  mutual  aid,  and  comfort,  man  began  to  live 
in  towns  and  cities.  This  necessitated  supplying  water  for 
the  common  use  by  gravity  from  some  higher  source  to  foun- 
tains, by  means  of  hand  pumps  from  wells,  or  by  utilizing 
natural  springs.  These  town  fountains,  pumps,  or  springs  are 
still  prominent  objects  in  the  cities  of  the  Old  World,  and  also 


2  FUMPIXG  ^lACHINERY 

in  our  own  colonial  towns.  As  the  town  developed  the  supply 
was  carried  into  some  of  the  houses.  Finally,  as  the  burgher 
became  a  manufacturer  and  was  compelled  to  dig  into  the 
earth  for  his  raw  materials,  it  became  necessary  to  clear  his 
excavations  of  the  surface  and  subsurface  waters  which  filled 
them. 

Modern  civilization  has  demanded  more  and  more  as  pump- 
ing machinery  has  been  perfected,  until  to-day  running  water 
in  unlimited  supply  is  found  not  only  in  the  houses  of  cities, 
large  and  small,  but  even  the  isolated  farm  often  has  its  own 
water  works,  giving  ample  water  to  each  room  of  its  house, 
barn,  and  dairy.  The  latest  demand  is  the  supply,  at  a  moment's 
notice,  of  large  quantities  of  water,  under  great  pressure,  to 
the  congested  districts  of  trade  in  our  large  cities  for  fire 
protection. 

Thus  from  earliest  times  may  be  traced  a  demand  for  some 
means  of  raising  water  for  man;  for  his  herds  and  lands;  for 
the  purpose  of  clearing  his  mines,  and  finally  for  his  own  per- 
sonal convenience  as  well  as  the  protection  of  his  property. 
The  pumping  machine  has  developed  from  a  very  crude  origin, 
it  is  true,  yet  its  earliest  types  were  so  effective  that  they  may 
be  found  in  use  to-day  although  in  modified  form. 

The  present  work  will  not  consider  the  sources  of  water 
supply,  the  gauging  of  the  flow,  the  distribution  of  water,  its 
analysis  or  purification,  or  any  of  the  problems  of  hydraulics 
save  those  which  are  concerned  with  the  raising  of  water  from 
one  level  to  another. 

Two  of  the  earliest  forms  of  pumps  are  the  Shadoof  and  the 
Noria,  the  former  being  common  in  Egypt,  while  the  latter  is 
found  in  China  and  along  the  Euphrates  as  well  as  on  the  Nile. 
The  first  of  these  is  the  ordinary  well  sweep  seen  on  many  old 
farms  in  this  country.  A  leather,  earthen,  or  woven  bucket 
is  attached  to  an  arm  by  means  of  ropes  or  tree  branches  and 
ropes.  This  arm  is  tied  to  a  crossbeam  supported  in  crotches 
of  tree  trunks  planted  in  the  ground  at  the  edge  of  some  river 
or  well.  The  arm  supporting  the  bucket  is  counterweighted 
by  a  stone  or  a  mud  ball,  so  that  there  will  be  practically  no 


HISTORICAL  DEVELOPMENT 


FIG.  i.— Shadoof. 


FIG.  2. — Shadoofs  in  Series. 


PUMPING  MACHINERY 


weight  to  lift.  A  man  then  pulls  on  the  bucket  support, 
putting  the  bucket  beneath  the  water,  and  then  allows  the 
counterweight  to  lift  it  to  the  proper  level,  where  he  empties 

the  water  into  the  canal  or 
basin.  From  the  canal  it 
flows  to  the  land  which  is 
to  be  irrigated.  At  times 
a  series  of  these  shadoofs  is 
placed  in  line  (Fig.  2),  each 
shadoof  raising  the  water  a 
portion  of  the  total  lift  which 
would  be  too  great  for  any 
one  machine.  The  shadoof 
is  probably  the  oldest  ap- 
paratus for  raising  water, 
although  the  simple  bucket 
attached  to  a  rope  (Fig.  3) 
must  have  been  used  in 
early  times.  Wilson  states 
that  in  India  one  or  two 
men  operate  a  Paecottah  (the 
southern  Indian  name  for 
the  Egyptian  shadoof),  and 
lift  from  1000  to  3000  cubic 
feet  in  from  six  to  eight 
hours.  The  lift  varies  from 


FIG.  3. — Bucket. 


5  to  12  feet.  This  apparatus  is  known  also  as  the  Swape 
and  in  upper  India  it  is  called  the  Ldt.  Buckley  states  that 
the  experiments  of  Wilson  showed  that  two  men  could  lift 
57,600  cubic  feet  through  a  height  of  one  foot  in  ten  hours, 
while  one  man  could  lift  33,000  cubic  feet. 

The  Noria  (Fig.  4)  is  a  machine  by  which  the  water  of  a 
stream  is  raised  in  buckets  attached  to  a  wheel,  the  wheel  being 
moved  by  the  stream  or  in  some  cases  by  animal  power.  The 
Chinese  claim  to  have  used  these  as  early  as  one  thousand  years 
before  the  Christian  Era.  One  of  them  is  described  as  con- 
sisting of  eighteen  or  twenty  arms  with  paddles,  to  the  periph- 


HISTORICAL   DEVELOPMENT  5 

ery  of  which  is  attached  a  number  of  buckets.  At  the  lowest 
limit  -of  the  motion  of  the  wheel  these  buckets  dip  below  the 
water  and  are  filled.  The  motion  of  the  wheel  thus  raises 
the  water  to  a  higher  level.  In  some  wheels  the  impact  of  the 
stream  on  the  buckets  is  sufficient  to  drive  the  wheel,  while 
in  other  cases  the  wheel  is  provided  with  additional  vanes  to 
drive  it.  A  simple  Chinese  wheel  (Fig.  4)  is  formed  of  bamboo. 
The  spokes  of  this  wheel  are  attached  to  a  central  shaft  and 
cross  each  other  at  two-thirds  the  distance  to  the  periphery, 


FIG.  4. — Noria. 

where  they  are  firmly  lashed  together,  while  at  the  end  they 
are  fastened  to  end  pieces.  These  act  as  the  buckets,  the 
natural  joint  in  the  bamboo  forming  the  bottom  of  the 
bucket.  The  triangle  formed  by  the  spokes  and  the  end 
piece  is  filled  with  bamboo  basket  work,  thus  forming  a 
paddle  to  drive  the  noria.  The  end  tubes  are  cut  and 
fastened  at  such  an  angle  that  when  the  vane  is  in  a  hori- 
zontal plane  the  end  tube  is  inclined  upward  at  an  angle  of 
twenty  degrees  so  that  the  buckets  will  not  discharge  their 
contents  until  they  pass  the  level  of  the  axle.  The  buckets 


6  PUMPING   MACHINERY 

finally  discharge  into  a  trough  which  conducts  the  water  to  a 
canal  or  reservoir  from  which  it  can  be  used.  The  wheels  are 
from  20  to  40  feet  in  diameter.  In  the  case  of  a  wheel  20  feet 
in  diameter,  containing  twenty  buckets  2  inches  in  diameter 
and  4  feet  long,  70,000  gallons  were  raised  in  twenty-four 
hours  when  the  wheel  made  four  revolutions  per  minute. 

These  wheels  were  used  throughout  the  East  in  Asia  and 
Egypt.  Colonel  Chesney  of  the  British  army  reports  several 
of  them  in  operation  along  the  Euphrates  as  motors  and  pumps. 
Some  of  these  were  arranged  so  that  their  axles  could  be  raised 
by  means  of  stones  in  order  that  their  depths  of  immersion 


FIG.  5. — Modern  Nona. 

would  be  the  same  at  different  stages  of  the  river.  Along  the 
river  aqueducts  are  seen  which  carried  the  water  away,  and 
walls  or  dams  are  found  which  served  to  raise  the  water  at  the 
season  of  small  flow  and  also  served  to  direct  the  flow  to 
the  openings  in  the  walls  where  the  norias  were  placed.  At 
the  center  of  the  river  openings  were  left  through  which  boats 
could  pass. 

An  American  wheel  of  modern  application  (Fig.  5)  shows 
that  the  idea' of  centuries  ago  is  still  of  value  as  the  basis  of  an 
irrigating  apparatus;  as  arranged  here  the  full  diameter  is 
almost  available  for  the  lift.  Such  a  machine  is  simple  and 
very  effective. 


HISTORICAL  DEVELOPMENT  7 

The  name  "noria"  is  also  applied  to  a  development  of  the 
older  ancient  machine,  in  which  the  buckets  are  attached  to  a 
chain  or  rope  which,  with  its  series  of  buckets  or  pitchers,  is 
placed  over  the  wheel  and  extends  down  to  the  water  at  con- 
siderable distance  below.  In  this  case,  however,  the  wheel  is 
driven  by  animal  power.  This  form  is  known  as  the  Persian 
Wheel  (Fig.  6),  while  according  to  Buckley,  the  term  Sakias 
is  used  for  it  in  certain  other  places.  To  operate  the  Persian 
wheel  a  buffalo  or  camel  is  driven  around  a  vertical  axis,  to  an 
arm  of  which  it  is  yoked.  The  axle  contains  a  cog  wheel 
(Figs.  6,  7,  and  8)  of  crude  form  which  engages  with  a  second 
cog  wheel  mounted  on  a  horizontal  axis  ori  which  is  placed  the 


FIG.  6. — Sakias. 


band  or  chain  wheel  for  the  support  of  the  bucket  chain.  The 
water  is  discharged  from  the  buckets  into  a  trough  and  flows 
to  whatever  fields  are  to  be  irrigated.  It  is  stated  that  two 
bullocks  will  lift  2000  'cubic  feet  of  water  per  day  and  the 
heights  of  the  lift  will  vary  from  25  to  100  feet.  The  apparatus 
is  common  in  the  East,  and  in  1890  the  American  consul  at 
Cairo  reported  20,000  of  these  in  use  in  the  upper  and  lower 
Nile  valleys. 

A  form  in  which  the  buckets  are  attached  to  the  wheel  as 
in  the  old  noria,  but  in  which  the  wheel  is  driven  by  animal 
power  as  in  the  Persian  wheel,  is  still  used  at  times,  and  is 
known  as  the  Sakias.  The  term  Taboot  is  applied  to  such  an 
apparatus  when  the  buckets  are  replaced  by  bags  made  of 
animal  skins, 


PUMPING  MACHINERY 


FIG.  7.— Persian  Wheel. 


FIG.  8.— Persian  Wheel. 


HISTORICAL  DEVELOPMENT 


I 


' 


I 

I 

1 

I 

I 


«s« 

FIG.  9.— Joseph's  Well. 


FIG.  10. — Chain  Pump. 


I, .  "i||IK,n  Illlll'    .ill.  "illllli t-nlllh  ,i!i.ol 


FIG.  n.— American  Form  of  Persian  Wheel. 


10 


PUMPING  MACHINERY 


-  An  old  Persian  wheel  whose  date  of  construction  is  not  known 
is  located  at  Joseph's  well  at  Cairo.  This  is  shown  in  Fig.  9. 
The  total  depth  is  almost  300  feet,  and  this  depth  is  divided 
into  two  lifts.  The  inclined  passageway  around  the  upper 
shaft  enables  the  animals  used  for  driving  to  be  taken  down 
to  the  lower  driving  wheel.  The  Romans  built  similar  appar- 
atus, calling  them  Roman  Buckets.  They  were  similar  to 
our  coal  and  grain  elevators.  Another  development  of  this 
same  apparatus  consisted  of  a  number  of  square  discs  mounted 
on  a  chain  and  fitted  into  a  square  pipe  like  small  pistons. 
These  chain  discs,  which  have  become  the  common  chain 


FIG.  12. — Mental. 

pump  of  to-day,  have  been  used  in  China  from  time  imme- 
morial. Fig.  10  shows  the  construction  of  this.  The  pallets  or 
chaplets  A  A  attached  to  the  chain  B  are  really  loose-fitting 
pistons  and  in  the  application  shown  there  is  no  suction. 

The  Persian  wheel  has  been  applied  in  America  (Fig.  n); 
with  it  a  horse  is  used  and  the  crude  gearing  and  buckets  are 
replaced  by  metal  parts.  The  principle,  however,  is  the  same 
as  that  of  the  old  wheel.  With  such  a  machine  a  horse  is  said 
to  lift  500  cubic  feet  per  hour  through  a  height  of  20  feet. 

The  Mental,  Katweh  or  Latha  is  a  form  of  apparatus 
used  in  Egypt  and  India  to-day  (Fig.  12).  It  consists  of  a 
basket  attached  to  two  ropes  in  such  a  manner  that  two  men 
may  swing  it  into  a  stream  when  swinging  it  in  one  direction, 


HISTORICAL    DEVELOPMENT 


11 


while  on  the  return  swing  the  basket  by  a  dextrous  twist  is 
discharged.  In  this  manner  two  men  may  raise  20,000  cubic 
feet  of  water  one  foot  in  a  day  of  ten  hours. 

The  Doon  (Fig.  13)  consists  of  a  long  trough  pivoted  near 
one  enoT  and  balanced  by  a  stone  attached  to  an  overhead  lever. 
The  operator  stands  on  the  long  counterbalanced  end  of  the 
trough,  overcoming  the  counterbalance  and  causing  the  outer 
end  of .  the  trough  to  dip  into  the  water.  On  stepping  from 
this,  the  weight  lifts  the  trough  and  discharges  the  water  into 
the  ditch  above  the  stream  or  pond. 

Another  primitive  apparatus  was  the  Mot,  which  consisted 


FIG.  13. — Doon. 

of  a  bag  of  skin  attached  to  a  rope.  The  bag  was  raised  to 
the  surface  by  oxen,  discharged,  and  then  dropped  back  for 
another  supply.  This  apparatus  is  effective,  as  two  bullocks 
and  a  man  can  lift  79,000  cubic  feet  one  foot  in  ten  hours. 

The  double  Zig-Zag  Balance  (Fig.  14)  is  shown  by  Mr. 
H.  M.  Wilson,  as  an  apparatus  used  in  Asia  Minor  and  Egypt. 
Two  men  oscillate  the  frame  and  thus  cause  the  water  to  flow 
past  wooden  valves  at  the  intersection  of  the  steps,  and  the 
water  gradually  passes  from  one  step  to  the  next.  On  each 
swing,  water  is  lifted  and  gradually  travels  upward  to  the  dis- 
cliarging  trough. 


12 


PUMPING   MACHINERY 


Fig.  15  illustrates  a  method  shown  on  certain  Egyptian 
monuments.  This  should  hardly  be  called  a  pumping  appli- 
ance, but  it  serves  to  illustrate  the  great  importance  of  irriga- 
tion when  in  ancient  times  such  expensive  methods  were 
employed. 

This  brings  one  near  the  beginning  of  the  Christian  Era, 
and  at  this  time  there  were  several  important  applications  of 


FIG.  14. — Zig-zag  Balance. 

mechanical  principles  to  the  raising  of  water.  The  use  of 
suction  to  raise  water  was  applied  and  valves  were  added  to 
tubes  carrying  water  while  the  movable  diaphragm,  partition 
or  piston  closely  fitting  the  water  vessel  or  tube  was  employed. 
The  principle  of  the  syphon  was  known  at  this  time,  and 
the  syringe,  which  employed  these  various  principles,  was  in 
use. 

The  principle  of  atmospheric  pressure  was  not  understood, 
although  it  was  employed  in  the  suction  of  water  in  early 


HISTORICAL   DEVELOPMENT 


13 


machines.     It  was  not  until  the  time  of  Torricelli,  1644,  that 
this  was  fully  comprehended. 

Fig.   16    is   known    as  a   Tympanum.      It   was    employed 


FIG.  15. — Egyptian  Irrigation. 

in  Egypt.     A  spiral  tube  is  attached  to  the  face  of  the  wheel, 
which  is  driven  by  the  current.     The  end  of  the  tube  dips 


FIG.  16. — Tympanum. 

beneath  the  stream,  and  as  the  wheel  turns  part  of  this  water 
is  caught  within  the  tube  and  is  gradually  lifted  to  the  hub 
of  the  wheel,  where  it  is  discharged  into  an  irrigation  flume. 


14 


PUMPIXG   MACHINERY 


The  Archimedean  Screw  (Fig.  17)  is  one  of  the  first  com- 
binations of  a  movable  diaphragm  in  a  tube.  By  turning  the 
screw  it  is  seen  that  the  water  is  compelled  to  travel  upward. 
The  helical  surface  A  on  the  axis  B  is  rotated  only,  but  the 
effect  of  it  is  that  of  axial  movement. 

The  application  of  the  principle  of  suction  and  one  of  the 
first  uses  of  heat  to  lift  water  is  described  by  Hero  of  Alexandria 
(dr.  120  B.C.)  in  his  "Pneumatics."  In  this  apparatus  the 
heat  from  the  burning  sacrifice  on  the  altar  A  (Fig.  18)  is  used 
to  open  the  doors  B  of  the  temple.  This  was  accomplished 


FIG.  17. — Archimedean  Screw. 

by  the  pressure  produced  on  heating  the  air  contained  within 
the  hollow  altar  A.  The  air  was  led  into  the  top  of  the  sphere 
C  and  its  pressure  drove  the  water  into  the  bucket  Z),  the  weight 
of  which  acted  on  ropes  attached  to  the  trunnions  E  of  the 
doors.  When  the  sacrifice  was  burned  the  air  contracted, 
producing  a  vacuum  in  C;  the  water  was  sucked  back  from  D 
and  the  doors  were  closed  by  the  weight  F.  In  this  apparatus 
it  is  important  to  notice  the  use  of  pressure  to  force  water 
from  one  vessel  to  another,  and  of  a  vacuum  or  suction  to 
draw  water  into  a  vessel. 

From  the  description  of  Hero  it  is  not  possible  to  know 


HISTORICAL   DEVELOPMENT 


15 


whether  he  invented  any  of  the  devices  described,  but  it  is 
reasonable  to  suppose  that  many  of  them  were  the  inventions 
of  Ctesibius,  who  made  several  mechanical  inventions.  The 


FIG.  18. — Temple  Pump. 


FIG.  19. — Force  Pump. 


FIG.  20. — Sun  Fountain. 


Force  Pump  is  ascribed  to  him  and  to  Hero.  This  pump 
(Fig.  19)  may  have  been  suggested  by  the  noria,  in  which 
valves  were  used  at  the  bottom  of  the  buckets  to  facilitate 


16 


PUMPING  MACHINERY 


filling.  There  were  two  single-acting  pistons,  moved  up  and 
down  by  a  cross  rod  or  beam,  and  although  the  pumps  were 
single  acting  the  stream  was  continuous,  owing  to  the  arrange- 
ment of  the  pumps.  It  is  to  be  noted  that  this  was  intended 
for  a  fire  pump,  and  is  the  earliest  fire  engine  of  which  there  is 
record. 

Another    machine    (Fig.   20)   called   a   Fountain  uses   the 
expansive  force  of  heated  air  to  operate  it      Water  is  driven 


FIG.  21. — Steam  Pump. 


FIG.  22. — Fountain  of  de  Caus. 


from  the  vessel  A  through  the  syphon  B  by  the  expansion  of 
the  air  in  the  top  of  A  when  exposed  to  the  sun's  rays.  This 
water  then  falls  into  the  box  D  and  when  the  vessel  A  cools 
water  is  drawn  up  from  D  through  the  pipe  E.  Although 
Hero  describes  it  in  this  manner,  it  is  well  to  note  that  some 
form  of  valve  must  be  placed  in  C  and  E  to  have  the  action 
proceed  as  described. 

The   use   of   steam   pressure   was   suggested   by   Giovanni 
Baptista  della  Porta  in  1601  in  an  apparatus '  (Fig.  21)  in  which 


HISTORICAL  DEVELOPMENT 


17 


steam  is  generated  in  a  boiler  A ,  from  which  it  is  discharged 
into  the  top  of  a  vessel  B  filled  with  water.  A  pipe  C  reaches 
to  the  bottom  of  B,  and  when  the  steam  pressure  is  exerted 
on  the  water  in  B  it  is  forced  from  the  discharge  pipe  C.  The 
apparatus  is  the  same  as  that  of  Hero,  and  della  Porta  suggests 
in  his  book,  published  in  1601,  that  the  vacuum  caused  by  the 
condensation  of  the  steam  be  used  for  filling  the  vessel  with 
water.  Della  Porta  should  have  shown  another  pipe  leading 
from  the  vessel  B  to  the  source  of  water  supply.  It  is  interest- 
ing that  della  Porta  suggested  the  separation  of  the  boiler  and 


FIG.  23. — Pump  of  da  Vinci. 

the  pump  cylinder,  a  very  valuable  point,  although  in  1615 
Solomon  de  Caus  proposed  to  combine  the  two  by  placing'  the 
equivalent  of  the  vessel  B  on  the  fire  and  heating  the  water  in 
it  for  the  purpose  of  driving  the  same  by  the  pressure  of 
the  steam.  (Fig.  22.) 

Before  this  time,  in  the  fifteenth  century,  Leonardo  da  Vinci 
suggested  the  pump  shown  in  Fig.  23.  The  lantern  wheel  A 
was  turned  by  two  cranks  and  this  motion  was  transmitted 
to  a  cylinder  B  through  the  toothed  wheel  C,  and  a  helical 
groove  in  the  cylinder  B  caused  a  piston  rod  D  to  travel  back 
and  forth  by  means  of  a  pin  on  the  rod  which  fitted  into 


18  PUMPING   MACHINERY 

the  helical  groove.  The  movement  of  the  solid  piston  of  the 
cylinder  E  would  then  suck  water  from  the  well  through  the 
suction  pipe  F  and  discharge  it  through  the  discharge  pipe 
G.  Valves  must  have  been  employed  here,  although  not  shown, 
and  the  machine  at  once  suggests  the  complete  understanding 
of  the  action  of  the  piston  within  a  cylinder. 

In  1630  a  patent  was  granted  to  David  Ramsey e  by  the 
English  king  which  covered  two  points:  "  to  raise  water  from 
low  pitts  by  fire,"  and  "  to  raise  water  from  low  places  and 
mynes  and  coal  pits,  by  a  new  waie  never  yet  in  use." 

There  is  no  record  of  what  he  did,  but  this  seems  to  be  one 
of  the  first  useful  applications  of  heat  to  the  raising  of  water. 

While  the  use  of  heat  for  pumping  was  progressing,  the 
application  of  water  power  and  animal  power  was  being  devel- 
oped extensively.  In  England  there  was  the  installation  of 
the  London  Bridge  Water  Works  in  I58i;  by  which  the  current 
of  the  Thames  was  used,  while  in  France  Agostino  Ramelli 
published  a  book  describing  his  inventions,  including  pumps. 
This  book  was  published  in  1588,  and  the  pumps  described 
were  operated  by  men  and  animals  as  well  as  by  the  currents 
of  streams.  According  to  an  old  account  the  application  of 
the  lift  pump  was  made  in  1581,  "  when  Peter  Morrys  was 
given  a  grant  by  the  Lord  Mayor  and  Commonalty  of  the  city 
of  London  for  the  term  of  500  years  for  supplying  and  con- 
veyance of  water  into  houses  by  pipes  from  an  artificial  force 
from  London  Bridge  on  condition  that  he  pay  ten  shillings 
annually  into  the  chamber  of  London."  He  was  authorized 
to  use  the  first  arch  of  London  Bridge  for  this  purpose. 

In  a  paper  before  the  American  Water  WTorks  Association, 
Mr.  T.  W.  Yardley  quotes  from  a  description  published  in  1633 
as  follows:  "  The  present  supply  of  good  water  for  London  is 
like  to  be  very  much  enlarged  by  the  great  improvement  of 
the  water  works  of  Peter  Morrys  before  mentioned,  who,  being 
a  Dutchman,  in  the  twenty- third  year  of  Queen  Elizabeth, 
first  gave  assurance  of  his  skill  in  raising  Thames  water  so 
high  as  should  supply  the  upper  parts  of  London;  for  the  Mayor 
and  Aldermen  came  down  to  observe  the  experiment^  and 


HISTORICAL   DEVELAPMENT 


19 


they  saw  him  throw  water  over  St.  Magnus  steeple,  before  which 
time  no  such  thing  was  known  in  England  as  the  raising  of 
water."  The  success  of  this  invention  was  such  that  Morrys 
obtained  additional  grants  for  two  other  arches  under  London 
Bridge.  *  The  second  grant  was  for  2000  years  and  was  finally 
secured  by  the  New  River  Water  Company. 

In  1731  Henry  Height  on,  the  engineer,  described  the  London 
Bridge  water  works  in  the  "  Philosophical  Transactions,"  and 
accompanied  his  minute  account  by  an  engraving.  It  may  be 
that  the  pumping  apparatus  had  changed1  from  the  time  of  its 
first  installation,  but  that  is  not  known. 

There  were  three  water  wheels  (Fig.  24.)  at  the  time  of  this 


FIG.  24. — London  Bridge  Water  Works. 

description,  each  about  20  feet  in  diameter.  These  wheels  A 
were  carried  on  heavy  frames  BB  and  were  raised  and  lowered 
with  the  tide.  The  frames  B  supporting  the  wheel  A  were 
pivoted  at  the  axles  of  the  lantern  wheels  CC,  so  that  although 
the  axle  of  the  wheel  A  was  raised  and  lowered  the  pin  wheels 
DD  attached  to  A  on  its  axle  were  always  in  contact  with 
the  lantern  wheels  CC.  The  beams  BB  were  supported  at 
their  outer  ends  by  chains,  which  were  raised  or  lowered  by  the 
cranks  EE  acting  on  a  series  of  crown  wheels  and  lantern 
wheels. 

The  axles  of  the  lantern  wheels  CC  were  connected  by  a 


20  PUMPING  MACHINERY 

coupling  to  crank  shafts  FF,  each  provided  with  four  cranks, 
which  in  turn  were  connected  to  a  series  of  vibrating  levers. 
These  levers  were  connected  to  pump  rods  at  each  end,  although 
in  the  figure  one  set  of  pumps  is  omitted  for  clearness.  There 
were  sixteen  pumps  to  the  wheel  with  cranks  arranged  so  that 
four  of  them  would  work  alternately.  The  pump  cylinders 
were  4  feet  9  inches  long  and  7  inches  in  diameter;  they  were 
fixed  to  the  top  of  an  iron  cistern  which  contained  the  proper 
foot  valves,  while  the  discharge  valves  were  placed  in  another 
box. 

According  to  the  description  of  this  pump  the  gears  were 
so  arranged  that  the  crank  turned  2,\  times  per  turn  of  the 
main  wheel.  With  the  several  wheels  used  in  1733,  containing 
52  forcers  or  cylinders  in  all,  1954  hogsheads  were  pumped 
per  hour  while  the  main  wheels  made  six  turns  per  minute. 
This  would  occur  with  full  tide,  and  at  that  rate  46,896  hogs- 
heads would  be  pumped  per  twenty-four  hours. 

Beighton  states  that  the  amount  of  leakage  or  "  slip  "  in 
the  pump  amounted  to  from  20  to  25  per  cent  of  its  displace- 
ment. The  reasons  for  this  he  gives:  ist,  the  amount  of 
valve  lift  will  cause  leakage;  2d,  no  leather  can  be  made 
strong  enough  for  pistons.  He  states  that  loose  leathers  cause 
leakage  while  tight  leathers  cause  excessive  friction. 

Such  pumps  excite  admiration  when  the  amount  of  experi- 
ence possessed  at  that  day  and  the  state  of  the  art  of  both 
machinists  and  millwrights  is  considered. 

While  the  works  of  Peter  Morrys  were  being  constructed 
in  England,  a  book  appeared  in  Paris,  1588,  by  Captain  Agos- 
tino  Ramelli,  an  Italian  engineer.  An  account  of  this  book  on 
the  various  machines  of  Ramelli  is  given  by  Mr.  W.  F.  Durfee 
in  Gassier' s  Magazine  for  June,  1895.  Among  the  machines 
for  pumping  water  given  by  Mr.  Durfee  three  have  been  taken 
to  illustrate  methods  described  by  Ramelli,  and  also  to  illus- 
trate the  state  of  the  art  at  this  time.  A  number  of  the  machines 
show  that  Ramelli  was  familiar  with  the  action  of  the  piston 
and  the  rotary  engine  as  well  as  with  the  principles  of  gearing 
to  transmit  power  between  shafts  at  an  angle  and  not  in  the 


HISTORICAL   DEVELOPMENT  21 

same  plane;     and   for  the  purpose  of  obtaining  reciprocating 
motion  from  continuous  rotary  motion. 

In  Fig.  25  a  method  is  shown  for  draining  an  excavation 
or  site  beside  a  flowing  stream.  The  wheel  A  is  driven  by  the 
stream.  Attached  to  its  axle  or  shaft  B  is  a  series  of  cranks 
moving  a  set  of  levers  CC,  which  cause  the  two  shafts  D  to 


FIG.  25. — Ramelli's  Pump. 

oscillate.  Each  of  the  shafts  DD  has  four  arms  which  serve 
to  drive  the  pump  rods  of  four  submerged  pumps. 

It  is  to  be  noted  that  in  both  this  machine  and  that  of  Peter 
Morrys  the  crank  and  connecting  rod  are  in  evidence  long 
before  the  time  of  James  Watt.  Here  are  also  seen  successful 
methods  of  securing  reciprocating  motion  from  rotary  motion. 

The  wheel  A  of  Fig.  26  was  turned  by  a  man  walking  on  the 
inside,  and  in  this  the  crank  B  gave  a  reciprocating  motion  to 


22 


P  UM  PI  NG   MA  CHI  N  ER  Y 


FIG.  27. — Ramelli's  Rotary  Pump. 


HISTORICAL   DEVELOPMENT 


23 


the  chain  CC,  which  pulled  in  each  direction  on  the  arm  E, 
driving  the  arms  DD  up  and  down  and  pumping  on  each  stroke, 
one  pump  F  forcing  water  on  the  up  stroke  of  E  and  the  other 
on  the  down  stroke. 

This  not  only  shows  the  application  of  the  crank  and  con- 
necting rod,  but  there  is  much  ingenuity  displayed  in  arranging 
a  chain  to  operate  on  both  strokes. 

The  rotary  pump  (Fig.  27 )  of  Ramelli  is  deserving  of  partic- 
ular notice  as  a  pump  and  also  for  the  ingenious  arrangement 
of  its  gears.  For  some  reason  the  chain  was  not  carried  down 
to  the  axle  of  the  pump.  It  may  have  been  to  keep  this  out 
of  the  water  or  for  the  purpose  of  getting  a  higher  speed.  The 
pin  and  lantern  wheels  at  A  and  the  crown  and  lantern  wheels 
at  B  serve  to  increase  the  number  of 
rotations,  while  the  spiral  gears  at  C 
serve  only  to  change  the  direction  of 
motion. 

In  1628,  two  years  before  Ramseye 
secured  his  patent  from  King  Charles 
for  the  use  of  fire,  Edward  Somerset, 
Marquis  of  Worcester,  is  thought  to 
have  installed  an  apparatus  for  raising 
water  at  Raglan  Castle.  He  did  not 
secure  a  patent  on  this  until  1663,  how- 
ever, and  there  were  no  drawings  nor 
even  a  mcdel  with  his  patent.  From 
the  description  of  the  apparatus  in  his 
patent  and  from  the  grooves  in  the 
walls  of  Raglan  Castle,  an  idea  of  the 
construction  and  operation  of  the  ma- 
chine may  be  formed,  although  it  is  not  FlG  28._WorCester»s  Pump. 
certain  that  this  is  correct  in  every  detail. 

Two  vessels  A  A  (Fig.  28)  are  connected  to  a  boiler  through 
a  valve  B.  In  the  figure  the  connection  to  the  boiler  is  made 
to  the  right-hand  vessel  and  the  steam  from  the  boiler  presses 
on  top  of  the  water  in  A  and  forces  it  out  through  the  valve  C. 
During  this  action  the  right-hand  foot  valv  •»  D  is  held  down  on 


24  PUMPING  MACHINERY 

its  seat.  The  water  is  driven  to  a  height  corresponding  to  the 
steam  pressure,  and  for  that  reason  high  steam  pressure  was 
required  for  great  heads.  While  the  right-hand  vessel  A  is 
discharging,  the  condensation  of  steam  in  the  left-hand  vessel 
produces  a  vacuum  in  the  vessel  and  water  is  drawn  up  through 
the  foot  valve  D. 

On  the  reversal  of  the  valves  B  and  C  the  operation  just 
described  is  repeated  with  opposite  vessels  and  so  the  operation 
became  continuous. 

Worcester  introduced  this  for  the  water  supply  at  Vauxhall 
near  the  old  city  of  London.  Upon  examination  it  is  seen  that 
apart  from  the  useful  application  of  the  invention  there  is 
nothing  new  in  this  apparatus,  as  the  elements  are  all  found  in 
della  Porta's  work. 

In  1683  Sir  Samuel  Moreland,  the  master  mechanic  of  the 
laboratory  of  Charles  II,  published  a  book  from  Paris  on  "  The 
Elevation  of  Water  by  All  Sorts  of  Machines."  He  had  been 
sent  to  Paris  by  Charles  on  business  relating  to  the  water  works 
which  the  king  had  erected.  In  his  book  Moreland  speaks 
of  the  expansion  of  steam  and  its  pressure,  showing  that  a  good 
idea  of  the  pressure  and  volume  of  saturated  steam  was  common 
in  that  day.  He  also  refers  to  the  duty  of  his  pumping  engines, 
using  the  term  as  it  is  used  to-day — the  amount  of  work  per 
hundred  weight  of  coal.  The  question  of  the  application  of 
steam  to  the  raising  of  water  was  one  which  not  only  occupied 
Moreland,  but  many  other  mechanicians,  on  account  of  the 
difficulty  which  was  then  experienced  in  clearing  the  shafts 
of  the  English  mines  from  the  vast  quantities  of  water  which 
collected  therein  from  the  underground  flow.  Many  mines 
had  to  be  abandoned  because  of  the  cost  of  carrying  the  mines 
deeper  when  this  expense  of  draining  was  too  great.  The 
work  was  mostly  done  by  animal  power  and  the  cost  was  rather 
startling  when  compared  with  the  steam  pump  of  even  the 
early  days. 

It  is  to  be  noted  that  Moreland  introduced  and  invented 
the  plunger  type  of  pump  in  1675.  This  type  of  pump  was 
one  in  which  an  enlarged  end  of  a  rod  was  forced  into  a  cham- 


HISTORICAL   DEVELOPMENT 


25 


her,  displacing  the  water.  This  was  followed  shortly  afterward 
by  a  bucket  pump,  in  which  the  water  passed  through  valves 
in  the  piston  on  the  down  stroke,  as  was  the  case  in  the  later 
engines  of  Simpson  used  at  Thames-Ditton. 

In  1698  Thomas  Savery  patented  the  design  of  an  engine 


FIG.  29. — Savery 's  Pump  of  1702. 

for  freeing  the  mines  of  Cornwall  from  water.  It  was  the  first 
steam  apparatus  applied  to  this  kind  of  work.  In  1699  he 
submitted  a  model  to  the  Royal  Society  of  London  and  success- 
ful experiments  were  made  with  it.  A  model  fire  engine  was 
exbihited  before  King  William  III  at  Hampton  Court  in  1698, 
and  the  success  of  this  led  to  the  granting  of  the  patent  which 


26  PUMPING  MACHINERY 

read:  "  A  grant  to -Thomas  Savery,  Gent'L,  of  the  sole  exercise 
of  a  new  invention  by  him  invented  for  raising  of  water  and 
occasioning  motion  to  all  sorts  of  mill  works  by  the  impellant 
force  of  fire,  which  will  be  of  great  use  for  draining  mines, 
serving  towTis  with  water  and  for  the  working  of  all  sorts  of 
mills,  when  they  have  not.  the  benefit  of  water  nor  constant 
winds;  to  hold  for  14  years  with  usual  clauses." 

The  apparatus  is  almost  identical  with  that  of  Worcester, 
and  it  is  not  known  whether  or  not  Savery  knew  of  the  earlier 
work.  The  form  of  his  pump  of  1702,  which  is  an  improvement 
on  that  of  1698,  is  shown  in  Fig.  29.  It  consists  of  a  main 
boiler  A  and  the  pumping  chambers  B  and  C.  The  water  from 
the  tank  G  discharges  through  the  valve  /  on  one  of  the  pumping 
vessels,  which  condenses  the  steam  in  that  vessel  and  the  vacuum 
produced  thereby  draws  water  through  the  suction  pipe  L 
and  foot  valve  E.  When  steam  is  then  admitted  from  A  through 
the  steam  valve,  the  water  is  forced  out  through  the  valve  D 
into  the  discharge  pipe.  \Vhen  the  water  is  low  in  the  boiler  A, 
the  auxiliary  boiler  is  filled  from  the  main  K  through  the  pipe 
H  and  valve  /.  This  water  is  driven  into  the  main  boiler  by 
raising  steam  in  M.  The  pipe  connecting  the  two  boilers  is 
carried  to  the  bottom  of  the  auxiliary  boiler  M  and  the  steam 
pressure  on  the  water  drives  it  over  into  the  main  boiler  A, 
provided  the  pressure  in  M  is  higher  than  that  in  A . 

Savery  seems  to  have  been  a  wide-awake  promoter  and 
advertiser,  for  he  began  a  systematic  scheme  for  making  his 
invention  known.  He  explained  it  to  the  Royal  Society  and 
presented  them  with  a  drawing  and  specifications  which  ap- 
peared in  their  "  Transactions,"  and  he  published  a  prospectus, 
called  "  The  Miner's  Friend;  or  a  Description  of  an  Engine 
to  raise  water  by  fire  described  and  the  manner  of  fixing  it  in 
mines,  with  an  account  of  several  uses  it  is  applicable  to,  and 
an  answer  to  the  objections  against  it.  London,  1702."  This 
invention  of  Savery  was  intended  to  do  away  with  the  great 
expense  in-  the  use  of  animal  power  to  operate  the  pumps  of 
'the  mines,  which  not  only  were  expensive,  but  reached  their 
limit  of  capacity,  so  that  workings  could  not  be  carried  much 


HISTORICAL   DEVELOPMENT 


27 


farther.     In  one  mine  500  horses  were  employed  in  handling 
the  water. 

Savery's  improvements  were  the  addition  of  surface  con- 
densation, the  secondary  boiler,  and  the  use  of  water  cocks 
It  must  be  remembered  that  this  machine  could  not  be  used  in 
deep  mines,  as  sufficient  steam  pressure  could  not  be  carried. 
The  joints  of  the  sheets  forming  the  boilers,  pump  chambers, 
and  other  parts  were  fastened  by  solder,  and  at  high  tempera- 


FIG.  30. — Piston  Pump  of  Papin. 

tures  these  joints  would  not  hold.  Desaguliers  reports  such 
accidents. 

The  Savery  engine  was  built  by  others  after  his  time — by 
such  engineers  as  Desaguliers-  and  Gravesande. 

While  Savery  was  using  the  direct  action  of  steam  for  raising 
water,  Papin,  in  Germany,  was  proposing  the  use  of  the  piston 
to  separate  the  steam  and  the  water,  thus  leading  to  the  further 
application  of  two  pistons  of  different  sizes,  so  that  the  steam 
pressure  did  not  have  to  be  equal  to  the  water  pressure.  In 
1707  Papin  proposed  a  water  pump  shown  in  Fig.  30.  This 
was  brought  out  in  Cassel,  Germany.  In  this  machine  steam  was 
generated  in  the  boiler  A  and  was  conducted  through  the  cock 
B  to  the  cylinder  C.  It  acted  on  the  piston  H  and  forced  water 


28  PUMPING   MACHINERY 

through  the  discharge  valve  G  into  an  air  chamber,  from  which 
it  was  delivered.  Water  was  admitted  through  the  valve  K 
after  the  piston  had  been  driven  to  the  end  of  its  stroke.  Papin 
realized  the  cooling  effect  of  the  cylinder  walls  and  suggested 
the  introduction  of  a  piece  of  heated  iron  E  to  warm  up  the 
piston  before  the  introduction  of  the  steam.  The  weighted 
cover  D,  similar  to  Papin's  safety  valve,  served  for  the  intro- 
duction of  the  iron. 

Although  Papin  had  in  his  pump  the  possibility  of  making 
the  steam  pressure  different  from  that  of  the  water,  it  was 
not  to  him  'that  the  honor  for  the  first  use  of  pistons  of  various 
sizes  is  .due.  Pistons  for  the  forcing  of  water  were  explained  by 
Ramelli  and  Leonardo  da  Vinci,  but  these  pistons  were  driven 
through  some  mechanical  medium  from  water  power  or  man 
power,  and  it  was  Thomas  Newcomen,  a  blacksmith  of  Dart- 
mouth, England,  who  first  employed  steam  for  the  operation 
of  a  water  piston  of  different  size  from  the  steam  piston.  This 
at  once  enabled  mines  to  be  sunk  to  a  greater  depth,  as  pump- 
ing could  now  be  done  with  greater  efficiency.  The  Savery 
engine  was  used  to  lift  water  350  feet,  but  with  the  engine 
invented  by  Newcomen  the  height  was  limited  only  by  the 
strength  of  the  materials  employed. 

The  engine  of  1705  is  shown  in  Fig.  31.  Steam  is  generated 
in  boiler  A  at  about  atmospheric  pressure,  and  as  the  piston  of 
the  cylinder  B  is  drawn  up  by  the  unbalanced  weight  of  the 
pump  rod  C  steam  is  drawn  into  the  cylinder  from  the  boiler 
through  a  valve  at  the  top.  When  the  piston  reaches  the  top 
of  its  stroke  the  valve  is  closed  and  water  is  sprayed  into  the 
steam  space  from  reservoir  D,  thus  condensing  the  steam  in 
the  cylinder  and  producing  a  vacuum.  The  air  pressure  on 
top  of  the  piston  is  then  sufficient  to  force  the  piston  down, 
raising  the  pump  rod  C,  and  with  it  the  pump  piston  at  its 
lower  end.  The  water  of  condensation  then  falls  through  a 
pipe  G  into  a  hot  well,  the  height  of  the  water  in  the  drain  being 
fixed  by  the  vacuum  in  the  cylinder.  The  walking-beam  E 
served  to  connect  the  two  piston  rods  by  chains  and  sectors, 
and  the  rod  C  could  be  of  any  length,  as  the  greater  part  of  its 


HISTORICAL   DEVELOPMENT 


29 


weight  and  that  of  the  column  of  water  in  the  discharge  pipe 
could  be  balanced  by  counterweights,  on  a  rod  such  as  F  on 
either  side  of  the  center.  Sufficient  weight  was  left  unbalanced 
to  cause  the  piston  of  cylinder  B  to  rise  when  low-pressure 
steam  was  admitted. 

Since  the  engine  was  really  driven  by  atmospheric  pressure 
and  was  operated  by  steam  at  practically  no  pressure  above  the 
atmosphere,  it  was  known  as  "the  atmospheric  engine." 


FIG.  31. — The  Atmospheric  Engine  of  Newcomen. 

Indeed,  so  low  was  the  steam  pressure  that  the  weight  of  the 
safety  valve  H  was  sufficient  to  keep  it  closed. 

The  try  cocks  K  served  to  indicate  the  quantity  of  water 
in  the  boiler,  and  the  cock  above  the  cylinder  B  was  used  to 
introduce  water  on  the  top  of  the  piston  to  keep  the  edge  of  it 
airtight.  The  use  of  a  spray  inside  the  cylinder  to  condense 
the  steam  resulted  from  the  discovery  that  leakage  of  water 
around  the  piston  condensed  the  steam  more  quickly  and 


30  '  PUMPING   MACHINERY 

better  than  the  original  method  of  cooling  the  cylinder  wall 
by  spraying  it  with  water. 

It  is  remarkable  that  a  man  of  such  little  training  as  New- 
comen  appears  to  have  had  should  have  been  able  to  combine 
the  necessary  elements  to  make  such  a  great  machine.  He  did 
not  occupy  a  very  high  position  in  the  town;  however,  he  was  a 
good  workman.  When  he  and  his  colleague,  John  Galley, 
wrote  to  Dr.  Hooke,  the  famous  physicist,  in  regard  to  their 
use  of  a  steam  cylinder  and  piston  to  drive  a  separate  pump, 
lie  advised  against  their  plan.  They  were  not  to  be  put  down, 
however,  and  in  1705  they  secured  a  patent. 

The  engine  was  then  applied  to  drain  mines  and  pump 
water,  but  Newcomen  and  Galley  did  not  have  sufficient  mathe- 
matical knowledge  to  properly  design  their  machines  and  they 
had  many  failures,  while  their  successes  were  accidents. 

In  1713  Humphrey  Potter,  a  boy  who  operated  valves  by 
hand,  arranged  an  automatic  method  of  shutting  off  the  steam 
and  water  by  the  use  of  beams,  strings,  and  catches,  and  made 
the  machine  independent  of  an  attendant.  Henry  Beighton 
improved  this  in  1718  by  substituting  a  vertical  beam  with 
pins  which  struck  the  valve  handle  as  it  was  raised  and  low- 
ered by  the  walking  beam.  The  vertical  beam  was  known  as 
a  "  plug  rod,"  "  plug  tree,"  or  "  plug  frame." 

These  pumps  were  built  for  various  purposes  and  were  of 
different  dimensions;  one  in  1714  for  Ansthorpe  in  Yorkshire 
vras  23  inches  in  diameter  and  of  a  6-foot  stroke.  It  made  15 
strokes  per  minute.  One,  described  by  Farey,  was  8  inches  in 
diameter  on  the  water  side  and  24  inches  on  the  steam  side. 
The  stroke  was  60  inches  and  there  were  15  strokes  per  minute. 
This'  pump  lifted  water  162  feet  and  the  water  column  on  the 
piston  weighed  over  3500  pounds,  which  with  660  pounds  of 
unbalanced  weight  of  the  pump  rod  necessitated  almost  5000 
pounds  on  the  piston.  Such  a  pressure  could  be  obtained 
wiVh  a  vacuum  of  21  inches  of  mercury.  The  pump  devel- 
oped 8  horse  power.  Another  engine  at  Griff  in  Warwick- 
shire cost  ^"150  per  year  to  operate  it  and  displaced  500  horses 
at  an  expense  of  £900  per  year.  The  first  Newcomen  engine 


HISTORICAL  DEVELOPMENT  31 

was  introduced  on  the  continent  in  1723  at  Konigsberg,  Hun- 
gary. In  1735  cast  iron  was  used  in  place  of  wrought  iron  for 
the  parts  of  the  engine. 

The  engine  was  improved  by  many  engineers.  About  1769 
John  Smeaton,  one  of  the  most  distinguished  engineers  of  his 
day,  built  several  engines  with  greater  stiokes  than  those  usu- 
ally employed.  By  using  the  proper  diameters  for  his  pistons 
he  was  enabled  to  get  much  higher  speed.  Before  building 
pumps  he  experimentally  determined  the  proper  proportions 
of  the  engine  and  so  improved  its  construction. 

Before  the  last  quarter  of  the  century  these  engines  were 
introduced  to  such  an  extent  that  the  coal  mines  of  Coventry 
and  Newcastle,  the  tin  and  copper  mines  of  Cornwall,  blowing 
engines  of  the  English  and  Scotch  furnaces,  the  docks  of  Cron- 
stadt  in  Russia,  the  lowlands  of  Holland  and  the  salt  mines 
of  Hungary  bore  testimony  to  the  success  of  this  invention. 
The  mines  were  carried  to  greater  depths,  the  cost  of  pumping 
water  and  air  was  reduced,  and  the  supply  of  water  to  towns 
was  more  certain. 

One  of  Smeaton's  Newcomen  engines  is  shown  in  Fig.  32. 
The  figure  shows  the  cylinder  A  connected  with  the  boiler  by 
means  of  the  steam  pipe  B.  The  boiler  is  placed  in  another 
building.  The  valve  C  admits  steam  to  the  cylinder  through 
the  admission  pipe,  which  is  carried  above  the  bottom  of  the 
cylinder  so  as  to  keep  the  injection  water  from  entering  it. 
When  the  steam  is  admitted,  the  piston  is  driven  or  pulled 
upward  and  when  the  top  of  its  stroke  is  reached  the  upward 
movement  of  the  plug  tree  or  working  plug  D  acts  on  the 
handles  E  through  pins,  turning  the  axle  F,  and  the  Y  or 
"  tumbling  bob  "  G  is  thus  moved,  shifting  the  rod  H  and 
handle  /  and  thus  shutting  off  the  steam.  At  the  same  time 
the  handle  K  opens  the  valve  M ,  allowing  water  from  the  cis- 
tern N  to  enter  the  cylinder  through  the  spray  head  0.  This 
immediately  condenses  the  steam  in  the  cylinder  and  the 
vacuum  produced  permits  the  atmospheric  pressure  to  drive 
down  the  piston.  The  pins  P  and  the  springs  Q  stop  the  down- 
ward motion  at  the  proper  point.  At  this  lowest  point  the 


32 


PUMPING  MACHINERY 


FIG.  32. — Newcomen  Engine  of  Smeaton. 


HISTORICAL   DEVELOPMENT  33 

drain  pipe  R  is  opened,  allowing  the  condensed  steam  and  con- 
densing water  to  discharge  into  the  hot  well  5.  The  sniffing 
valve  T  is  then  opened,  allowing  any  air  to  escape.  The  feed 
water  of  the  boiler  is  taken  from  the  hot  well.  The  cistern 
N  is  supplied  by  a  jack-head  pump  £7,  driven  from  a  small 
beam.  Both  beams  get  their  motion  through  chains  and 
sectors,  so  that  there  is  always  a  straight  pull  on  the  piston  or 
pump  rods.  The  cock  V  admits  water  around  the  piston  so 
that  the  oakum  packing  of  the  piston  is  kept  in  proper  con- 
dition. The  excess  of  this  water  is  carried  off  through  the  drain 
pipe  to  the  hot  well.  The  main  beam  is  carried  on  sectors  so 
as  to- reduce  the  friction. 

To  improve  the  efficiency,  Smeaton  covered  the  steam 
side  of  the  piston  with  planks  and  when  the  injection  water 
contained  salts  which  would  form  a  scale,  the  water  in  the  , 
hot  well  was  not  used  for  boiler  feed,  but  the  clear  feed 
water  was  passed  through  a  coil  of  pipe  immersed  in  the  hot 
well. 

The  next  important  step  in  the  improvement  of  the  steam 
engine  and  pump  was  that  of  James  Watt.  The  invention  of 
this  man  was  one  of  the  greatest  events  in  the  history  of  civili- 
zation, as  it  not  only  improved  the  existing  machines,  but  in 
his  specifications  are  contained  the  fundamental  ideas  of  all 
modern  improvements  to  the  steam  engine.  This  event  clearly 
demonstrates  what  may  be  done  by  a  careful  and  detailed 
study  of  existing  conditions. 

The  early  history  of  Watt,  who  was  born  in  1736,  is  one  with 
which  every  engineer  should  be  familiar.  After  many  trials 
and  successes  in  the  south  of  England  he  came  back  to  Glasgow 
and  was  employed  to  repair  some  of  the  apparatus  belonging  to 
the  university.  In  1763  he  repaired  its  model  of  the  New- 
comen  engine,  and  this  led  to  his  making  a  study  of  the  history 
of  the  steam  engine.  He  read  the  treatise  of  Desaguliers  and 
the  works  of  others.  In  this  study  he  learned  of  the  accom- 
plishment of  Savery,  Newcomen,  and  those  who  had  preceded 
them.  Watt  now  began  a  series  of  experiments  on  the  action 
of  the  engine,  determining  quantitative  relations  between  the 


34  PUMPING  MACHINERY 

amounts  of  steam,  cooling  water,  and  the  heat  of  steam  and 
water.  He  discovered  the  great  loss  by  radiation  and  absorption 
and  was  led  to  use  non-conducting  materials  for  his  vessels  as  well 
as  for  the  coverings  of  them.  Not  having  experimental  data 
for  this  work,  he  made  a  series  of  original  experiments  on  the 
temperature  and  pressure  of  steam  at  whatever  points  he  could 
observe  these  and  constructed  a  curve  to  give  values  at  other 
points.  He  determined  the  amount  of  steam  used  by  the 
Newcomen  engine  and  the  amount  which  should  have  been 
used  had  none  of  the  steam  condensed;  in  addition  he  com- 
pared the  amount  of  injection  water  used  in  the  engine  with 
the  amount  which  should  have  been  used.  These  experiments 
showed  him  that  three-fourths  of  the  steam  taken  into  the 
cylinder  was  wasted  and  that  the  engine  used  four  times  as 
much  injection  water  as  it  should  have  used.  His  calculation 
showed  him  at  once  that  the  method  of  producing  the  vacuum 
was  a  poor  one,  as  the  cylinder  had  to  be  heated  by  steam  at 
each  stroke  so  that.it  could  be  filled,  and  then  this  heat  was 
removed  again  on  the  condensation  of  the  steam.  He  then 
tried  to  keep  the  cylinder  hot,  which  necessitated  that  the 
steam  be  taken  from  it  for  condensation.  He  invented  the 
independent  condenser,  which  made  his  improvement  complete. 
After  constructing  a  number  of  experimental  machines  and 
after  many  vicissitudes  he  took  out  his  patent  in  1769  in  con- 
nection with  Dr.  Roebuck. 

The  patent  of  1769  gave  the  following  description: 
"  My  method  of  lessening  the  consumption  of  steam,  and 
consequently  fuel,   in  fire  engines,   consists  in  the  following 
principles: 

"  ist.  That  the  vessel  in  which  the  powers  of  steam  are  to 
be  employed  to  work  the  engine,  which  is  called  '  the  cylinder ' 
in  common  fire  engines,  and  which  I  call  '  the  steam  vessel ' — 
must,  during  the  whole  time  that  the  engine  is  at  work,  be 
kept  as  hot  as  the  steam  which  enters  it;  first,  by  inclosing 
it  in  a  case  of  wood,  or  any  other  materials  that  transmit  heat 
slowly;  secondly,  by  surrounding  it  with  steam  or  other  heated 
bodies;  and  thirdly,  by  suffering  neither  water  nor  other  sub- 


HISTORICAL  DEVELOPMENT  35 

stances  colder  than  the  steam  to  enter  or  touch  it  during  that 
time. 

"  2dly.  In  engines  that  are  to  be  worked,  wholly  or  partially 
by  condensation  of  steam,  the  steam  is  to  be  condensed  in 
vessels  distinct  from  the  steam  vessel  or  cylinder,  though 
occasionally  communicating  with  them.  These  vessels  I  call 
condensers;  and  while  the  engines  are  working,  those  condensers 
ought  at  least  to  be  kept  as  cold  as  the  air  in  the  neighbor- 
hood of  the  engines,  by  application  of  water  or  other  cold 
bodies. 

"  3dly.  Whatever  air  or  other  elastic  vapor  is  not  condensed 
by  the  cold  of  the  condenser,  and  may  impede  the  working  of 
the  engine,  is  to  be  drawn  out  of  the  steam  vessels  or  condensers 
by  means  of  pumps,  wrought  by  engines  themselves,  or  other- 
wise. 

"  4thly.  I  intend  in  many  cases  to  employ  the  expansive 
force  of  steam  to  press  on  the  pistons  or  whatever  may  be  used 
instead  of  them,  in  the  same  manner  as  the  pressure  of  the 
atmosphere  is  now  employed  in  common  fire  engines.  In  cases 
where  cold  water  cannot  be  had  in  plenty,  the  engines  may 
be  wrought  by  this  force  of  steam  only,  by  discharging  the 
steam  into  the  open  air  after  it  has  done  its  office. 

"  5thly.  Where  motions  round  an  axis  are  required,  I  make 
the  steam  vessels  in  form  of  hollow  rings  or  circular  channels, 
with  proper  inlets  and  outlets  for  the  steam,  mounted  on  hori- 
zontal axles  like  the  wheels  of  a  water  mill.  Within  them  are 
placed  a  number  of  valves  that  suffer  any  body  to  go  round  the 
channel  in  one  direction  only.  In  these  steam  vessels  are 
placed  weights,  so  fitted  to  them  as  to  fill  up  a  part  or  portion 
of  their  channels,  yet  rendered  capable  of  moving  freely  in 
them  by  the  means  hereinafter  mentioned  or  specified.  When 
the  steam  is  admitted  in  these  engines  between  these  weights 
and  valves,  it  acts  equally  on  both,  so  as  to  raise  the  weight  on 
one  side  of  the  wheel,  and  by  the  reaction  of  the  valves  suc- 
cessively, to  give  a  circular  motion  to  the  wheel,  the  valves 
opening  in  the  direction  in  which  the  weights  are  pressed,  but 
not  in  the  contrary.  As  the  vessel  moves  round,  it  is  supplied 


36  PUMPING  MACHINERY 

with  steam  from  the  boiler,  and  that  which  has  performed  ats 
office  may  either  be  discharged  by  means  of  condensers,  or  into 
the  open  air. 

"  6thly.  I  intend  in  some  cases  to  apply  a  degree  of  cold 
not  capable  of  reducing  the  steam  to  water,  but  of  contracting 
it  considerably,  •  so  that  the  engines  shall  be  worked  by  the 
alternate  expansion  and  contraction  of  the  steam. 

"  Lastly,  instead  of  using  water  to  render  the  piston  or 
other  parts  of  the  engine  air-  or  steam-tight,  I  employ  oils, 
wax,  resinous  bodies,  fat  of  animals,  quicksilver,  and  other 
metals  in  their  fluid  state." 

It  is  to  be  noted  that  these  claims  covered  the  following 
points : 

ist.  Lagging  and  jackets. 

2d.    Condensers. 

3d.   Air  pumps. 

4th.  Expansive  use  of  steam  and  the  non-condensing  engine. 

5th.  A  rotary  engine. 

6th.  Packings. 

Mathew  Boulton  became  the  partner  of  James  Watt,  and 
it  is  to  him  that  much  of  the  credit  of  the  actual  engine  is  due. 
He  was  the  owner  of  one  of  the  most  famous  manufactories 
of  the  day  at  Soho,  near  Birmingham,  England.  Here  he 
manufactured  ornamental  metal  ware,  gold-  and  silver-plated 
ware  and  works  of  art,  such  as  vases,  statues,  and  bronzes. 
His  factories  were  noted  for  the  good  work  done,  and  for  the 
broad  policy  of  management. 

Although  the  arrangement  of  the  partnership  was  agreed 
on  in  1769,  it  was  not  until  the  spring  of  1774  that  Watt  could 
go  to  Birmingham.  By  November  of  that  year  their  first 
engine  was  built.  The  form  of  this  is  shown  in  Fig.  33. 

Steam  enters  from  the  boiler  by  the  pipe  A  and  the  valve 
B  passing  to  the  steam  jacket  C.  The  condenser  D  is  then 
connected  to  the  cylinder  by  the  valve  E,  and  the  vacuum 
thus  produced  in  the  space  F  causes  the  piston  G  to  move 
downward,  and  steam  flows  in  above  .the  piston.  When  the 
piston  reaches  the  lower  end  of  the  stroke,  the  valves  B  and  E 


HISTORICAL  DEVELOPMENT 


37 


are  closed  while  the  valve  H  opens.  This  connects  the  spaces 
on  each  side  of  the  piston,  and  the  weights  of  the  pump  rods 
7  and  /  on  the  outer  end  of  the  beam  K  overbalance  the  weight 
of  the  piston  G  and  its  rod,  and  so  the  piston  is  pulled  rapidly 


FIG.  33. — Watt's  Engine. 

upward  to  the  top  of  the  cylinder,  the  steam  above  the  piston 
passing  over  to  the  lower  side. 

After  closing  H  and  opening  B  and  E  the  operation  is 
repeated,  and  the  air  pump  L  removes  the  condensed  steam 
and  the  air  from  the  surface  condenser.  The  pump  M  supplies 
the  cooling  water,  and  the  pump  N  takes  water  from  the  hot 
well  0  and  feeds  the  boiler.  The  pump  rod  of  the  air  pump 
contains  the  pins  which  operate  the  handles  of  the  valves. 


38  PUMPING  MACHINERY 

An  outlet  P  is  used  when  the  air  is  driven  out  from  the  cylinder 
and  the  air  pump  before  the  engine  is  started. 

There  was  much  trouble  in  getting  men  and  machines  to 
make  these  engines,  and  it  may  be  said  that  the  demand  for 
better  work  developed  the  machinist's  trade  of  that  day.  Much 
of  the  development  was  made  by  the  firm  of  Boulton  &  Watt. 

In  the  building,  erection,  and  operation  of  their  engines, 
Boulton  &  Watt  were  led  to  take  out  patents  for  the  following 
articles: 

A  letter  copy  press. 

A  cloth  dryer  by  the  use  of  steam  in  copper  rolls. 

Five  devices  for  getting  rotary  motion  from  reciprocating 
motion  without  the  use  of  a  crank. 

The  expansive  use  of  steam. 

The  double-acting  engine. 

The  double-coupled  engine. 

A  rotary  engine. 

A  trunk  engine. 

A  steam  hammer. 

Parallel  motion. 

The  engine  governor. 

Mercury  steam  gauge. 

Water  gauge. 

Steam-engine  indicator. 

Watt  seemed  to  be  one  who  could  always  find  some 
means  of  meeting  every  need:  when  it  took  too  much  time  in 
copying  his  reports  to  Boulton,  he  invented  the  copy  press; 
when  it  was  necessary  to  study  the  action  of  steam  in  the 
cylinder,  he  brought  out  his  indicator.  These  numerous  inven- 
tions do  not  indicate  that  the  firm  was  always  successful. 
Many  times  they  were  on  the  verge  of  bankruptcy,  and  had 
their  patent  not  been  extended  for  twenty-four  years,  when  it 
first  expired,  their  labor  would  have  been  in  vain,  because  of 
financial  straits.  The  extension  gave  them  the  needed  relief, 
and  at  the  expiration  of  the  patent  the  firm  was  in  good 
condition.  The  story  of  the  trials  and  successes  of  this  firm 
in  the  development  of  the  engine  is  given  in  the  biographies 


HISTORICAL   DEVELOPMENT 


39 


of  these  two  men,  and  the  student  is  recommended  to  study 
these  most  interesting  books. 

To  gauge  the  power  of  his  pumps,  Watt  introduced  the 
term  "  horse  power,"  in  so  common  use  to-day.  This,  with  the 
term,  "  duty,"  gave  those  using  pumps  a  method  of  comparing 


FIG.  34. — Hornblower's  Pump. 

the  operations  of  various  machines.  An  economical  operation 
was  aimed  at  in  all  of  this  work,  and  in  order  to  interest  the 
engine  attendants,  monthly  prizes  were  given  for  the  best 
results  in  duty  during  the  month.  An  operative  machine  had 
been  constructed,  and  it  was  now  their  object  to  improve  the 
efficiency  of  it. 

Boulton  died  in  1809  and  Watt  in  1819,  but  before  that 


40  PUMPING   MACHINERY 

time  these  men  had  given  the  business  over  to  their  sons. 
They  enjoyed  the  protection  of  their  fathers'  patents  until 
1801,  during  which  time  others  were  at  work  on  the  improve- 
ment of  the  engine  end  of  the  pump. 

Jonathan  Hornblower  patented  a  compound. engine  in  1781, 
although  Watt  claimed  this  invention.  The  engine  is  shown 
in  Fig.  34.  Steam  is  admitted  into  the  cylinder  A  from  the 
boiler,  and  from  this  cylinder  it  discharges  into  the  cylinder  B 
and  thence  into  the  condenser  C,  shown  in  section.  The  plug- 
tree  rod  D  serves  as  the  rod  for  the  air  pumps  as  well  as  to 
operate  the  valve  handles,  which  are  not  shown.  The  operation 
of  the  engine  is  practically  the  same  as  that  of  the  Watt  engine. 
To  start,  all  air  is  driven  out  by  allowing  boiler  steam  to  flow 
through  the  cylinders  and  condenser,  and  thence  through  the 
sniffing  valve  at  E.  The  condenser  will  now  condense  steam 
below  the  piston  in  B,  and  as  the  piston  descends  the  valve 
between  A  and  B  allows  steam  below  the  piston  of  A  to  expand 
and  press  down  on  top  of  the  piston  of  B.  The  steam  from  the 
boiler  enters  on  top  of  that  of  A.  When  the  bottom  of  the 
stroke  is  reached  the  boiler  and  condenser  are  cut  off  and 
the  top  of  each  cylinder  is  connected  to  the  lower  portion.  The 
weight  of  the  main  pump  rod  F  and  the  boiler  feed  pump  G 
pulls  the  pistons  upward  and  the  operation  is  repeated.  This 
was  declared  an  infringement  on  the  Watt  patent.  It  did  not 
give  a  much  higher  duty  than  the  best  single-cylinder  Watt 
engines  of  the  day,  although  the  same  idea  as  applied  by  Arthur 
Wolf  in  1804  with  higher  pressure  steam  gave  duties  of  from 
40,000,000  to  57,000,000  foot-pounds  per  bushel  of  coal,  while 
the  Watt  engine  gave  a  little  over  30,000,000. 

The  Bull  Cornish  pumping  engine  of  1798  was  brought  out 
by  William  Bull  and  Richard  Trevi thick.  This  type  of  engine 
is  the  one  which  remained  in  use  longer  than  any  other,  as  it 
was  much  simpler  than  that  of  Watt  and  had  all  of  the  ele- 
ments of  economy.  It  is  shown  in  Fig.  35. 

The  steam  cylinder  A  is  carried  on  the  timbers  BB,  extending 
from  the  walls  of  the  pump  house  in  such  a  manner  as  to  bring 
the  piston  rod  directly  over  the  pump  well.  The  piston  rod 


HISTORICAL   DEVELOPMENT 


41 


C  is  connected  to  the  pump  rod  Z),  and  this,  in  turn,  to  the 
counter  balancing  beam  E  by  the  rod  K' ' .  A  pump  rod  G  with 
its  pump  .F  is  also  shown.  The  counterweight  H  is  to  balance 
as  much  of  the  weight  as  is  thought  necessary.  The  rod  / 


FIG.  35. — Cornish  Pumping  Engine. 

actuates  the  piston  of  the  air  pump  and  is  used  as  a  plug  rod. 
The  valves  of  the  air  pump  are  in  the  base  and  the  piston  is 
solid.  The  tank  P  surrounded  the  air  pump  and  the  pipe  R; 
it  was  filled  with  water.  The  pipe  R  acted  as  a  condenser,  but 


42 


PUMPING  MACHINERY 


water  was  admitted  through  0,  making  it  really  a  jet  condenser. 
The  pins  'on  the  plug  rod  K  operated  the  rods  leading  to  the 
valves  at  L  and  M.  In  starting,  the  valves  were  operated 
from  the  floor  5  and  the  air  was  driven  out  from  the  cylinder, 
condenser,  and  air  pump  through  the  sniffing  valve  N,  which 
was  water-sealed. 

This  engine  was  adjudged  an  infringement  on  the  Watt 
patents,  which  prevented  its  introduction  for  some  time,  but 
it  was  afterward  used  exclusively  in  America  and  Europe 
with  many  improvements.  It  had  many  advantages  over  the 
other  engine  in  its  simplicity,  but  it  was  objectionable  in  that 
it  must  be  placed  directly  over  the  opening  of  the  mines.  It 
was  of  great  value  further  in  that  by  properly  selecting  the 


FIG.  36. — Action  of  Cornish  Engine. 

mass  of  the  parts  the  inertia  of  these  could  be  used  to  permit 
the  expansion  of  the  steam,  although  the  water  pressure  or 
resistance  was  constant.  This  may  be  shown  by  the  diagrams 
of  Fig.  36.  The  steam  pressure  in  excess  of  the  resistance  of 
the  water  is  used  in  accelerating  the  piston,  piston  rod,  pump 
rod,  balancing  beam,  and  counterbalance,  and  after  passing  the 
point  at  which  the  steam  pressure  equals  the  water  pressure 
the  inertia  of  the  parts  will  supply  the  deficiency  of  energy  on 
being  brought  to  rest  after  the  steam  pressure  falls  below  the 
resistance  of  the  water.  Neglecting  friction  the  area  abc  will 
equal  the  area  bde.  These  areas  will  always  be  a  measure 
of  the  energy  stored  up  in  the  moving  parts  and  will  be  a  func- 
tion of  the  maximum  velocity  and  the  mass  moved,  so  that 
by  changing  the  amount  of  mass  the  speed  of  the  apparatus 


HISTORICAL  DEVELOPMENT  43 

could  be  altered.  The  use  of  heavy  counterweights  at  times 
required  all  of  the  steam  pressure  on  the  up  stroke  of  the  weight 
to  move  them,  while  on  the  down  .stroke  the  excess  of  counter- 
weight acting  with  the  steam  was  used  to  lift  the  water  from 
deep  mines. 

It  is  important  to  note  the  advantage  of  use  of  steam  expan- 
sively, as  it  is  this  which  has  made  the  modern  pumping  engine 
with  a  fly  wheel  so  economical.  These  early  engines  were 
quite  efficient,  as  is  seen  from  the  table  below,  which  demon- 
strates the  great  advantage  of  this  Cornish  engine. 

DUTIES    IN    FOOT-POUNDS    PER    BUSHEL    (94    POUNDS)    OF 

WELSH    COAL 


In  1769,  the  Newc 
In  1772,  the  Newc 
In  1778  to  1815,  "V 
In  1820,  improved 
In  1826, 
In  1827, 
In  18-28, 
In  1830, 
In  1839, 
In  1850, 
In  1827,  highest  d 
In  1832, 
In  1842. 

omen  engine 

5,500,000  ft.-l 
9,500,000 
20,000,000 
28,000,000 
30,000,000 
32,000,000 
37,000,000 

43,350,000 

54,000,000 
60,000,000 
67,000,000 
97,000,000 
108.000.  ooo 

omen  engine,  improved  by  Smeaton 
V^att  engine 

Cornish  engine 

uty,  Consolidat< 
Fowey  Cons* 
United  Mine 

(averaged  

sd  Mines  

3ls  

s.  . 

This  engine  was  developed  into  the  beam  engine  and  was 
used  for  water  works.  Fig.  37  is  a  cut  of  a  Cornish  engine  of 
1840  for  the  East  London  Water  Works,  with  a  capacity  of 
6,000,000  gallons  per  twenty-four  hours.  This  Tepresented  the 
best  engine  of  the  day. 

While  the  Cornish  engine  was  being  used  for  pumping  water 
from  mines  and  for  the  water  supply  of  cities,  another  form  of 
pump  was  sucessfully  operated  in  1830  in  New  York  by  a  Mr. 
McCarty.  This  was  the  centrifugal  pump.  It  was  a  pump  which 
had  been  known  for  a  long  time,  as  Euler  discussed  its  theory 
in  a  paper  in  1754.  According  to  one  author  the  invention  of 
it  is  due  to  Denys  Papin  in  1689,  who  took  his  idea  from  Johann 
Jordan.  Jordan  designed  a  centrifugal  pump  in  1680.  Demour 
in  1730  invented  the  equivalent  of  a  centrifugal  pump.  It 


44 


PUMPING  MACHINERY 


consisted  of  a  tube,  Fig.  38,  mounted  on  a  vertical  axis  so  that 
the  lower  end  entered  the  water  to  be  raised  near  the  axis. 
On  turning  this  the  centrifugal  force  overcame  the  effect 'of 
gravity  and  the  water  rose.  In  1818,  a  few  years  before 


FIG.  37. — East  London  Water  Works. 

McCarty's  work,  a  centrifugal  pump  was  designed  in  Boston 
and  known  as  the  Massachusetts  pump.  It  was  a  successful 
machine.  Fig.  39  shows  the  general  arrangement  of  the  Boston 
pump.  The  vanes  were  parallel  to  radial  lines  and  removed 


HISTORICAL   DEVELOPMENT 


45 


several  inches  from  them.  They  were  placed  on  each  side 
of  a  disc,  and  this  runner  revolved  within  a  casing. 

After  McCarty  the  improve- 
ment of  this  form  of  pump  was 
undertaken  by  Blake  and  An- 
drews in  this  country  in  1831 
and  1839,  respectively,  and  by 
Appold,  Thompson,  and  Gwynne 
in  England.  The  original  blades 
of  the  Massachusetts  pump  were 
radial,  but  those  of  Andrews  in 
1846  were  curved,  as  shown  in 

Fig.    40.       This    pump    had    the    FlG>  38._Dera0ur's  Centrifugal  Tump. 

vanes  held  between  two  discs. 

This  patent  was  bought  by  John  Gwynne  for  England,  and  his 
firm  began  the  manufacture  of  these  pumps.  The  development 
of  the  centrifugal  pump  in  England  is  closely  connected  with 


FIG.  39. — The  Boston  Pump. 

the  history  of  this  firm.     They  were  the  constructors  of  the 
most  notable  installations  for  many  years. 

In  1848  Lloyd  took  out  a  patent  for  a  centrifugal  fan,  and 
Appold  began  the  manufacture  of  it  and  applied  it  to  the 


46 


PUMPING  MACHINERY 


lifting  of  water.  In  1851  he  exhibited  this  and  showed  its 
practicability.  The  tests  of  the  pump  with  the  curved  vanes 
showed  it  to  be  about  three  times  as  efficient  as  that  with 
straight  arms.  The  advantage  of  this  pump  is  its  ability  to 
lift  large  quantities  of  water  considering  the  space  occupied 
by  the  machine,  and  its  ability  to  pump  small  solid  particles 
without  clogging.  It  was  originally  thought  that  this  was  only 


FIG.  40. — Andrews  Pump. 


FIG.  41. — Serviere's  Rotary  Pump. 


applicable  to  low  lifts,  but  to-day  pumps  of  this  form  are  used 
for  lifts  of  several  hundred  feet. 

Another  old  form  of  pump  in  which  rotary  motion  of  the 
parts  is  utilized  is  shown  in  Fig.  41.  This  is  the  rotary  pump, 
and  is  an  old  invention  found  in  the  form  of  Fig.  41  among  a 
collection  of  models  made  by  Serviere,  a  Frenchman,  born  in 
I593-  This  is  one  of  the  best  forms  of  this  type,  as  will  be 
seen  later,  when  the  rotary  pump  will  be  examined  in  detail. 


HISTORICAL   DEVELOPMENT  47 

Since  the  action  of  this  pump  is  positive,  it  may  be  used  against 


FIG.  '42. — Ramelli's  Rotary  Pump. 

high  heads,  although  leakage  may  be  excessive,  due  to  the  wear 
which  occurs  in  the  parts. 
Some  claim  that  this  was 
the  invention  of  Pappen- 
heim,  a  German,  who  lived 
in  the  seventeenth  century. 
Ramelli,  whose  publica- 
tion of  1588  illustrates  a 
rotary  pump  shown  pre- 
viously in  Fig.  27,  uses  a 
slightly  different  form  from 
that  of  Serviere.  In  his 
pump,  Fig.  42,  a  series  of 
flat  pistons  are  driven  by 
a  rotating  cylinder  which  is 
placed  eccentrically  within 
the  outer  casing.  These 
flat  plates  are  held  out  by 

FIG.  43. — Sixteenth  Century  Rotary  Pumps. 

springs,  as   shown   in   the 

figure,  and  the  rotation  of  the  inner  cylinder  forces  the  watei 

through  the  machine, 


PUMPING    MACHINERY 


FIG.  44. — Watt's  Rotary  Pump. 


FIG.  45. — Eve's  Pump. 


HISTORICAL  DEVELOPMENT  49 

Another  form  of  rotary  pump  of  the  sixteenth  century  is 
given  in  Fig.  43.  As  seen  in  the  figure  the  space  A  is  being 
filled  through  an  opening  below  the  water  level  while  the  space 
B,  which  is  closed  by  the  sliding  partition  C,  is  being  discharged. 
The  sliding  partition  C  extends  from  one  side  of  the  casing  to 
the  other,  and  slides  through  the  stuffing  box.  After  rising 
to  the  highest  point  it  drops  by  its  weight,  which  is  made  suffi- 
cient to  overcome  the  friction  of  the  s tuning  box.  The  stuffing 
box  shows  the  form  used  at  that  day.  The  friction  of  this 
machine  was  very  great. 

A  form  of  rotative  pump  similar  to  that  patented  by  Watt 
in  1782  as  a  rotative  engine  is  shown  in  Fig.  44.  The  operation 
is  clear  from  the  figure;  the  flap  or  butment  A  serves  to  divide 
the  two  sides  of  the  pump  and  when  the  projecting  piece  or 
piston  strikes  the  butment  it  swings  on  its  pivot.  The  move- 
ment of  this  butment  is  controlled  by  a  heavy  spring,  or  by 
rods  and  cams,  so  that  it  is  held  against  the  water  pressure  from 
force  main,  moving  only  at  the  proper  time.  This  scheme  was 
altered  in  1825  by  J.  Eve  in  that  he  substituted  a  revolving 
cylinder  for  the  pivoted  butment  and  inserted  three  moving 
pistons  for  the  one.  The  small  drum  was  driven  by  gearing 
from  the  main  shaft  at  three  times  the  revolutions  of  the  main 
shaft,  as  it  was  one- third  the  size  of  the  main  drum.  This  is 
shown  in  Fig.  45. 

In  1805  John  Trotter  introduced  a  different  form,  Fig.  46, 
in  which  a  plate  was  driven  in  such  a  manner  as  to  touch  two 
fixed  concentric  drums,  its  position  being  radial.  The  operation 
of  the  machine  is  evident  from  the  figure.  There  is  some 
chance  for  leakage  after  the  piston  crosses  the  discharge  pipe 
and  before  it  crosses  the  suction  pipe,  so  that  it  is  really  neces- 
sary to  have  more  than  one  piston.  Fig.  47  illustrates  another 
form  of  this  used  for  water,  although  it  had  been  used  in  1790 
for  a  steam  engine. 

From  these  earlier  forms  a  number  of  new  rotaries  were 
developed  which  finally  became  a  variation  of  the  older  form 
of  Serviere,  as  will  be  seen  in  the  next  chapter. 

Another   old  type   to   be   mentioned  is   the   reciprocating 


50 


PUMPING  MACHINERY 


..FiG.  46. — Trotter's  Rotary  Pump. 


FIG.  47. — Four-Bladed  Rotary  Pump. 


HISTORICAL   DEVELOPMENT 


51 


rotary  form,  Fig.  48.  The  figure  shows  the  operation  of  the 
pump,  and  a  further  description  is  unnecessary.  The  objection 
to  these  and  the  rotary  pumps  is  the  fact  that  it  is  very  difficult 
to  keep  their  pistons  tight.  The  rotary  pumps  have  the  great 
advantage  that  the  flow 
of  water  is  always  in  the 
same  direction  through  the 
pump. 

A  pump  somewhat  al- 
lied to  the  rotary  is  the 
screw  pump,  Fig.  49.  The 
pump  illustrated  was  the 
invention  of  Revillion,  and 
was  patented  in  Paris  in 
1830.  It  consisted  of  a 
right-  and  a  left-handed 
screw  meshing  together, 
being  driven  in  opposite 
directions  at  the  proper 
speed  by  means  of  gears 
AB.  The  point  of  one  screw 
touches  the  root  of  the 
other,  and  thus  incloses  a 
definite  volume  of  water 
between  the  screw  and  the 
walls  of  the  pump  cham-  FlG"  48 -Reciproca^ng  Pump. 

ber,  which  travels  upward  as  the  screw  rotates.  At  the  end 
of  the  travel  this  water  is  forced  out  at  the  center. 

In  passing,  it  is  well  to  note  that  the  use  of  pumps  for  the 
extinguishing  of  fires  had  been  common  from  the  earliest  times, 
Fig.  19  being  one  described  by  Hero.  Until  about  1840,  how- 
ever, these  were  all  driven  by  hand  power.  Fig.  50  shows  a 
French  engine  of  1829.  This  was  hauled  to  the  fire  by  the  fire 
company,  which  in  America  was  a  very  important  social  organ- 
ization during  the  first  half  of  the  urban  history  of  the  last 
century. 

The  first  steam-driven  fire  engine  of  note  in  this  country 


52 


PUMPING  MACHINERY 


was  one  planned  by  Captain  John  Ericsson,  in  a  competition 
for  a  prize  offered  by  the  Mechanics  Institute  of  New  York, 

n 


FIG.  49. — Screw  Ptfmp  of  Revillion. 


FIG.  50. — French  Fire  Pump  of  1829. 

in  1840,  although  in  1830  Braithwaite  and  Ericsson  brought 
out  a  steam  fire  engine  in   London.    The  pump  developed 


HISTORICAL  DEVELOPMENT  53 

6  H.P.,  and  pumped  150  gallons  per  minute  a  distance  of  80 
or  go  feet.  It  was  drawn  by  horses  and  practically  was  the 
same  as  that  designed  by  Ericsson  for  New  York.  Before  this 
time  stationary  steam  fire"  pumps  were  used. 

Fig.  51  gives  a  view  of  the  Ericsson  engine.     The  boiler 
was  of  the  locomotive  type,  the  barrel  A  being  connected  to 


FIG.  51. — Steam  Fire  Pump  of  1840. 

the  firebox  B.  The  steam  pipe  C  supplied  the  cylinder  D 
with  steam.  The  water  cylinder  was  in  line  with  the  steam 
cylinder  and  above  it  was  placed  an  air  chamber.  The  smoke 
pipe  from  the  boiler  was  carried  around  the  air  cylinder  in 
the  form  of  a  serpent.  On  the  front  of  the  engine  was  a  blowing 
box  which  could  be  worked  by  the  cross-head  of  the  steam 
engine,  or  by  hand  or  by  a  crank  attached  to  the  wheels  of  the 
engine.  The  latter  arrangement  served  to  force  the  fire  within 
the  firebox  when  the  engine  was  on  its  way  to  a  conflagration. 
This  is  a  forerunner  of  the  modern  fire  engine,  and  it  is  note- 
worthy that  the  design  was  most  thoughtfully  and  carefully 
made. 

The  first  record  of  the  hydraulic  ram  was  that  of  Mr.  White- 
hurst  of  Derby,  England.  In  1772  he  erected  a  machine  shown 
in  Fig.  52.  A  was  a  spring  or 'reservoir  supplying  the  cock  C 


54 


PUMPING  MACHINERY 


through  the  pipe  B,  which  was  about  600  feet  long  and  H 
inches  in  diameter.  The  cock  was  16  feet  below  the  level  in 
A,  and  on  closing  this  after  drawing  water  the  momentum  of 
this  long  column  of  water  was  sufficient  to  force  the  water 


FIG.  52. — Hydraulic  Ram  of  Whitehursi. 

into  air  chamber  D,  which  was  under  the  pressure  of  the  higher 
reservoir  E. 

It   was   Montgolfier  in    1796   who  independently  invented 
the  same  scheme,  but  made  it  of  more  value  by  using  a  device 


FIG.  5*3. — Ram  cf  Montgclfkr. 

which  worked,  continuously  and  automatically,  in  place  of  the 
cock  C.  His  scheme  is  sho\vn  in  Fig.  53.  The  w.ater  descends 
from  the  source  of  supply  through  A,  escaping  at  B.  When 
the  water  has  acquired  a  certain  velocity  it  raises  the  ball  and 
closes  the  opening  at  B.  The  momentum  of  the  water  causes 
an  increase  of  pressure,  and  this  is  finally  sufficient  to  open  the 


HISTORICAL   DEVELOPMENT 


55 


valve  in  C  against  the  high  pressure  of  the  discharge.  The 
valve  at  B  may  be  of  the  disc  form,  but  opening  downward; 
the  principle,  however,  is  the  same  in  all  cases. 

The  development  of  the  ram  in  the  years  which  follow  the 
work  of  Montgolfier  consists  in   improvement   in  details,  the 


FIG.  54.— Flash  Wheels. 

latest  forms  of  the  present  day  being  quite  similar  to  this 
early  type. 

Scoop  wheels  or  flash  wheels,  Fig.  54,  were  used  from 
early  times.  They  were  in  reality  water  wheels  turning  back- 
ward. These,  as  will  be  seen,  have  been  used  to  some  advan- 
tage in  later  times.  They  were  used  extensively  in  Holland 
about  the  time  of  the  introduction  of  the  steam  engine. 


CHAPTER  II 
RECENT  HISTORY 

THE  year  1840  marks  an  era  in  the  history  of  pumping 
machinery,  for  it  was  in  that  year  that  Henry  R.  Worthington 
began  his  brilliant  inventions,  which  have  led  to  most  of  the 
modern  forms  of  steam  pumps. 

Before  this  lime  the  boiler-feed  pump  was  usually  driven 
by  the  main  engine  through  some  extension  from  the  piston 
or  pump  rod  or  by  an  auxiliary  rod  from  the  walking  beam. 
Worthington  was  concerned  in  the  design  of  a  steamboat  for 
canal  navigation.  It  happened  that  when  the  boat  was  stopped 
at  locks  or  by  obstructions  the  attendants  had  to  resort  to  the 
use  of  hand  pumps  to  keep  the  boiler  properly  filled.  An 
independent  steam  pump  was  thought  necessary,  and  on  Sep- 
tember 7,  1841,  the  pump  shown  in  Fig.  55  was  patented. 

In  the  illustration  this  pump  is  mounted  on  a  base  for 
exhibition  purposes,  although  it  was  originally  bolted  to  the 
side  of  the  boiler  setting  at  the  front.  Steam  enters  the  cylinder 
A  through  the  pipe  B,  the  steam  being  directed  to  either 
end  by  a  valve  in  the  steam  chest  C.  In  the  position  shown, 
the  valve  has  just  been  moved  so  as  to  admit  steam  to  the 
right-hand  end  of  the  cylinder.  This  drives  the  piston,  piston 
rod,  and  plunger  to  the  left.  The  plunger  D  drives  the  water 
frpm  the  cylinder  G  through  an  ordinary  conical  valve  in  the 
valve  box  E  to  the  feed  pipe  F  and  from  there  to  the  boiler. 
As  the  rod  moves  to  the  left  the  arm  H,  attached  to  the  rod, 
moves  the  tappet  rod  /  to  the  left.  Finally  near  the  end  of  the 
stroke  the  right  tappet  /  strikes  the  lever  K,  pivoted  at  its 
center,  and  forces  it  against  the  sloping 'top  of  the  rod  M.  This 
forces  the  rod  M  down  against  the  pressure  of  a  spring.  After 
the  lever  K  passes  over  the  point,  the  spring  pressure  suddenly 

56 


RECENT   HISTORY 


57 


forces  the  lever  K  over  to  the  left  away  from  the  tappet.  This 
moves  the  arm  N  to  the  right,  which  controls  the  steam  valve 
through  the  axis  of  the  arm  AT  so  as  to  admit  steam  to  the  left- 
hand  end  of  the  cylinder  A. 

The  piston,  piston  rod,  and  plunger  now  move  to  the  right, 


FIG.  55. — Original  Worthington  Pump. 

and  water  is  sucked  into  the  cylinder  G  through  a  conical  valve 
at  the  bottom  of  E  and  the  suction  pipe  0.  The  left-hand  tappet 
drives  the  lever  K  to  the  right  and  finally  the  spring  forces  it 
over,  suddenly  reversing  the  motion.  The  spring  action  was 
necessary  to  properly  reverse  the  steam  valve,  as  a  tappet 


FIG.  56. — Spring-Thrown  Valve. 

alone  would  permit  the  valve  to  cut  off  steam  from  each  end 
when  the  motion  was  slow.  It  is  necessary  to  have  the  valve 
reversed  positively  while  the  pump  has  positive  motion,  however 
slow. 

The  admission  of  steam  to  this  particular  pump  was  con- 
trolled by  a  float  in  the  boiler  so  that  when  the  water  became 


58 


PUMPING   MACHINERY 


low  the  pump  would  start  automatically  and  continue  until 
the  float  would  again  cut  oft  the  supply  of  steam.  The  pump 
shown  in  the  figure  was  in  service  for  twenty-five  years  and 
was  finally  bought  back  by  the  Worthingtons  as  a  cherished 
relic. 

In  1844  Worthington  used  a  helical  spring  in  a  casing  A, 
Fig.  56.  The  arm  B  pressed  against  a  helical  feather  or  pro- 
jection on  the  casing,  turning  the  casing  against  the  action  of 
the  spring.  When  the  arm  went  beyond  the  end  of  the  feather 
the  spring  forced  the  casing  over  suddenly,  thus  moving  the 
valve  at  C  by  turning  the  valve  rod,  moving  the  valve  across 
the  cylinder,  perpendicular  to  the  piston  motion. 

The  water  valves  at  D,  in  which  E  is  the  discharge  space 
and  F  the  suction,  are  of  the  forms  used  in  many  pumps,  of 


FIG.  57.— B  Valve. 

that  day.     The  pistons,  stuffing  boxes  and  cylindrical  parts  are 
quite  similar  to  those  of  the  present. 

The  B  valve,  Fig.  57,  was  invented  for  the  purpose  of 
admitting  steam  to  the  right-hand  end  of  the  cylinder  by  a 
motion  to  the  right  when  steam  is  above  the  valve.  In  the 
figure  it  is  seen  that  when  the  valve  is  moved  to  the  right, 
the  steam  above  the  valve  at  A  will  enter  the  space  B  through 
the  space  C  and  thus  pass  to  the  right-hand  end  of  the  cylinder 
through  D.  At  the  same  time  the  left-hand  end  E  is  con- 
nected to  the  exhaust  port  G  through  the  cavity  F.  This  is 
necessary  when  a  slide  valve  is  moved  directly  by  the  motion  of 
the  piston,  since  the  motion  to  the  right  moves  the  valve  to  the 
right  and  with  this  movement  steam  is  admitted  to  the  right- 
hand  end,  reversing  the  pump. 


RECENT  HISTORY 


59 


The  next  improvement  was  a  steam- thrown  valve,  Fig.  58. 
This  was  in  1849.  In  this  arrangement  an  auxiliary  valve 
rod  not  shown  moved  an  auxiliary  valve,  admitting  steam  to 
the  right  or  left  of  the  piston  A,  which  forced  the  small  cylinder 
to  the  right  or  left,  thus  moving  the  main  valve,  which  was 


FIG.  58.— Steam-Thrown  Valve. 

part  of  the  small  cylinder.  The  excessive  friction  from  the  steam 
on  top  of  the  auxiliary  piston  led  to  the  design  of  a  balanced 
valve  which  took  steam  in  through  the  ordinary  exhaust  pas- 
sage and  used  the  so-called  steam  chest  as  an  exhaust  chest.  . 


FIG.  59. — Positive  Water  and  Steam  Valve. 

Fig.  59  shows  a  pump  built  about  1850  in  which  both  the 
water  and  steam  valves  were  controlled  by  the  movement  of 
the  valve  rod.  This  was  in  a  measure  a  water  dash  pot  for 
the  purpose  of  stopping  the  motion  of  the  pistons.  Worthing- 
ton  used  this  in  some  cases,  but  he  did  not  advocate  its  general 
use. 


60  , 


PUMPING  MACHINERY 


In  1849  Worthington,  with  his  partner,  Wm.  H.  Baker, 
patented  a  relief- valve  motion.  This  scheme  is  shown  in 
Fig.  60,  where  radial  water  valves  or  clack  valves  are  used. 
The  invention  consisted  of  the  use  of  two  water  passages  A  A 
at  each  end  of  the  water  cylinder,  so  that  when  the  water 
piston  uncovered  the  inner  of  these  ports  the  pressure  in  front 
of  the  piston  was  relieved  suddenly  and  the  steam  in  the  steam 
cylinder  drove  the  piston  to  the  extreme  end,  moving  the  main 
slide  valve  B  over  by  means  of  the  arm  C  and  the  tappets  on 
the  valve  rod  D.  This  same  scheme  was  applied  by  cutting 
grooves  in  the  water-cylinder  bore  at  each  end. 

The  aim  in  all  of  these  later  pumps  was  to  simplify  the 


FIG.  60.— Relief  Valve  Motion. 

mechanism  so  that  the  pump  would  do  the  work  in  a 
positive  manner  and  still  be  simple  to  care  for  and  to 
operate. 

In  1850  Worthington  made  another  improvement  by  sub- 
stituting a  number  of  small  valves  for  the  four  large  valves 
used  on  pumps  heretofore,  and  also  employed  a  plunger  and 
ring  in  place  of  the  piston  of  former  pumps.  The  pump,  Fig.  61, 
was  used  on  the  steamer  "Washington."  There  were  thirty- 
six  of  these  small  valves  arranged  on  valve  decks  as  indicated 
in  the  figure.  The  large  number  of  -small  valves  gave  the 
requisite  amount  of  opening  with  a  small  lift  and  hence  the 
amount  of  leakage  passing  the  valves  when  the  pump  was 


RECENT  HISTORY 


61 


62 


PUMPING  MACHINERY 


reversed  was  greatly  reduced.  The  valves  were  of  rubber, 
one-half  inch  thick. 

The  arrangement  of  the  plunger  and  ring  of  Fig.  61  is 
markedly  different  from  the  pistons  of  the  former  figures,  and 
Worthington  adopted  it  for  several  reasons: 

ist.  It  gives  ample  space  above  and  below  for  the  valves. 

2d.  The  ample  space  around  the  plunger  forms  a  subsiding 
chamber  where  harmful  materials  may  settle  out  of  the  way 
of  the  plunger  packing. 

3d.  The  constant  protrusion  of  the  plunger  tends  to  carry 
foreign  matter  away  from  the  packed  joint. 


FIG.  62. — Steam  End  of  Savannah  Pump. 

4th.  The  construction  makes  the  removal  or  renovation  of 
the  parts  an  easy  matter. 

5th.  The  deflection  of  the  water  currents  is  less  than  in 
ordinary  arrangements.  The  friction  is  very  small. 

It  is  to  be  noted  in  passing  that  this  design  simplified  the 
construction  of  the  water  end  of  the  pump  as  far  as  the  foundry- 
work  and  the  machine-shop  work  were  concerned.  The  use  of 
a  solid  metal  packing  around  the  plunger  as  seen  at  A  was  an 
innovation,  but  it  has  proved  a  success.  It  was  made  long 
and  with  clear  water  the  wear  was  not  sufficient  to  cause  exces- 
sive leakage  for  some  time.  A  steam-thrown  valve  was  used 
in  this  case.  Worthington  was  not  satisfied  with  this  method 
of  valve  operation  and  after  his  invention  of  the  duplex  pump 
in  1859  ne  rarely  used  it,  although  inventors  such  as  Knowles, 
Blake,  and  others  continued  to  use  it. 


RECENT  HISTORY  63 

In  1854  Worthington  erected  his  first  water-works  pump  for 
trie  city  of  Savannah.  It  was  a  compound  pump  of  peculiar 
design,  in  that  the  high-pressure  cylinder  was  at  the  center  of 
the  annular  low-pressure  cylinder.  Another  new  feature  of  the 
pump  was  the  balancing  of  the  steam  valve  by  carrying  part 
of  the  pressure  on  the  back  of  the  valve  by  a  piston  supported 
from  the  steam-chest  cover  as  shown  in  Fig.  62.  This  figure 
illustrates  the  use  of  this  balancing  method  on  an  ordinary 
cylinder.  The  Savannah  engine  was  a  success  and  a  duplicate 
of  it  was  built  for  Cambridge,  Mass.,  in  1856.  This  engine, 
and  a  companion  engine,  built  soon  after,  were  of  the  following 
dimensions  : 

High-pressure  cylinder  .......................  12  inches  diameter 

Low-pressure  cylinder  .......................  25 

Plunger  ....................................  14  " 

Length  of  stroke  ..................  ..........  25  inches 

Capacity  ...................................  300,000  gals,  per  24  hrs. 

This  engine  was  the  best  of  its  day,  as  was  shown  by  tests 
made  for  the  Brooklyn  Water  Works  in  1857  and  1859,  when 
that  city  was  about  to  install  new  pumping  engines.  The 
results  of  these  tests  as  shown  in  the  report  of  Mr.  James  P. 
Kirkwood,  chief  engineer  of  the  Brooklyn  Water  Works,  were 
as  follows: 

Date  of  Test.  Name  of  Engine.  pe^lb^ 


April,  1857  Worthington  engine,  at  Cambridge  ..............  669,411 

June,  1857  ........  .  .....  675,746 

Jan.,  1857   Cornish  engine,  at  Jersey  City  .................  628,233 

July,  1857   Hartford  crank  engine  ................  ,  .......  646,994 

July,  1857  "         ......................  ..  614,426 

Jan.,  1860  Brooklyn  new  rotative  engine  .................  601,407 

June,  1856  Cornish  engine,  Philadelphia  ...................  589,903 

As  a  result  of  this  the  Worthington  pump  was  selected 
and  a  modification  of  the  Cambridge  pump  design,  known  as 
the  duplex  pump,  was  accepted;  this  was,  however,  abandoned 
on  account  of  trouble  with  the  contractor  and  it  was  not  until 
1863  at  Charlestown,  Mass.,  that  this  important  type  was 
installed  in  a  water  works.  It  will  be  remembered  that  the 
term  duty  as  used  by  Moreland  and  Watt  meant  the  useful 


64 


PUMPING   MACHINERY 


RECENT  HISTORY  65 

work  per  bushel  or  hundred  weight  of  coal,  but  in  later  years 
the  duty  was  figured  as  the  useful  work  per  1000  pounds  of 
steam  or  per  1,000,000  British  thermal  units. 

The  date  of  1859  is  that  of  the  invention  of  the  duplex 
pump,  one  of  the  simplest  contrivances  for  operating  the  valves 
of  a  pump,  doing  away  with  the  complex  arrangements  used 
heretofore  for  this  work.  This  then  became  for  many  years  the 
standard  type  of  pumping  engine,  as  the  Corliss  engine  became 
the  standard  mill  engine. 

To  make  clear  the  operation  of  the  duplex  pump  a  modern 
form  of  boiler-feed  pump,  Fig.  63,  is  illustrated.  This  con- 
sists really  of  two  pumps  placed  side  by  side.  In  this  figure 
the  piston  rod  A  of  the  front  pump  has  just  reached  the  right- 
hand  end  of  its  stroke,  and  the  bell-crank  lever  BC,  which 
extends  from  the  front  piston  rod  to  the  back  valve  rod  has 
moved  the  D  slide  valve  of  the  back  pump 'to  the  right,  uncover- 
ing the  steam  port  of  the  left-hand  end  of  the  rear  pump,  causing 
that  pump  to  move  to  the  right.  This  motion  is  transmitted 
to  the  front  valve  stem  through  the  reverse  lever  DE,  which 
connects  the  back  piston  rod  to  the  front  valve  rod.  The  D 
slide  valve  of  the  front  pump  is  moved  to  the  left,  admitting 
steam  to  the  right-hand  side  of  the  cylinder  and  thus  the  piston 
of  the  front  pump  moves  to  the  left.  This  motion,  in  turn, 
causes  the  back  pump  to  move  to  the  left  and  this  then  moves 
the  front  valve  so  that  the  piston  moves  to  the  right,  when  the 
operation  is  repeated. 

The  action  of  the  pump  may  be  represented  by  the  diagram, 
Fig.  64,  in  which  vertical  distance  represents  time  and  horizontal 
distance  represents  motion  of  the  pump.  The  solid  line  repre- 
sents the  front  pump  and  the  dotted  one  the  rear  pump.  At 
the  end  of  each  stroke  there  is  a  period  of  rest  while  the  other 
pump  follows  the  stroke  of  the  first  one.  This  is  accomplished, 
as  will  be  explained  later,  by  the  method  of  moving  the  valve 
by  the  valve  rod  or  its  equivalent. 

Not  only  is  the  claim  .for  a  simpler  valve  gearing  made 

or  this  pump,  but  there  should  also  be  a  steadier  discharge 

of  water,  because  as  one  pump  nears  the  end  of  its  stroke  the 


<56 


PUMPING   MACHINERY 


other  one  discharges  water  to  keep  up  the  flow  while  the  first 
reverses. 

This  duplex  pump  was  one  in  which,  by  the  use  of  an  addi- 
tional unit,  a  simple  form  of  valve  gear  was  obtained.  While 
Worthington  was  inventing  steam-thrown  valves  of  various 
forms  previous  to  his  invention  of  the  duplex  pump  the  same 


Stroke 
FIG.  04. — Action  of  Duplex  Pump. 

problems  were  being  investigated  by  others  in  this  country 
as  well  as  abroad.  It  is  not  the  intention  of  this  book  to  follow 
all  of  the  different  forms  of  simplex  pumps  invented  from  the 
time  of  Worthington  to  the  present,  as  his  inventions  mark 
the  era.  Later  a  number  of  modern  simplex  mechanisms  will 
be  examined  in  detail  to  illustrate  what  had  been  done  and 


RECENT  HISTORY  67 

what  forms  have  survived.  It  is  necessary  at  this  point, 
however,  to  mention  such  names  as  Blake,  Cameron,  Knowles, 
Earle,  Cope,  Maxwell,  Marsh,  Silver,  Dean,  Davison,  and 
Gordon  in  America;  and  Moreland,  Thompson,  Baummans, 
Tyler,  Clarkson,  and  Davies  in  Europe,  as  some  of  those  who 
had  been  working  on  this  problem.  Their  work  is  of  interest 
in  comparing  the  complicated  manner  of  the  earlier  forms  with 
the  simpler  forms  of  to-day. 

The  introduction  of  the  duplex  pump,  at  least  while  the 
patent  lasted,  caused  considerable  argument  as  to  the  relative 
merits  of  the  simplex  and  duplex  forms,  the  simplex  manufac- 
turers claiming  a  lack  of  positive  length  of  stroke  for  the  duplex 
against  a  full  stroke  of  the  simplex  to  offset  the  greater  com- 
plication. The  proper  adjustment  of  each  pump,  however, 
will  give  a  proper  stroke. 

Pumps  of  the  duplex  type  were  introduced  into  water 
works  in  1860,  and  in  1863  one  of  5,000,000  gallons  per  twenty- 
four  hours  was  installed  by  Worthington  at  Charlestown,  Mass., 
and  in  1871  a  larger  one  for  19,000,000  gallons  was  built  for 
the  city  of  Philadelphia.  The  earlier  engines  were  single 
expansion  on  the  steam  end. 

The  pumping  engine  of  Mr.  George  Shields  for  the  city  of 
Cincinnati,  which  was  built  in  1861,  was  one  of  the  largest  of 
its  day;  it  was  to  lift  9,600,000  gallons  per  twenty-four  hours 
against  a  head  of  170  feet.  It  was  direct  acting  and  had  no 
balance  beam.  The  steam  cylinder  was  100  inches  in  diameter 
and  had  a  stroke  of  12  feet.  The  pump  cylinder  was  45  inches 
in  diameter.  This  engine  was  one  of  the  unfortunate  structures 
of  American  engineering  in  that  it  cost  the  municipality  many 
times  its  original  estimated  figure;  but  to  its  credit  it  may  be 
said  that  twice  it  was  the  means  of  saving  the  city  from  a 
water  famine,  when  the  other  pumps  gave  out.  The  use  of 
vertical  engines  in  which  the  steam  and  water  cylinders  were 
placed  over  each  other  was  quite  common,  as,  for  example, 
in  the  Bull  Cornish  engines,  but  when  it  was  desired  to  eliminate 
the  inertia  bob  weight  to  cut  down  the  weight  of  the  engine, 
fly  wheels  were  introduced.  The  fly  wheels  were  driven  from  a 


68 


P  UMPI XC  MA  GHINER  Y 


beam  in  most  cases,  but  in  1868  Richard  Moreland,  Jr.,  and 
David  Thompson  invented  the  direct-acting  engine  shown  in 
Fig.  65.  In  this  the  cross-head  was  connected  to  the  pump 
plunger  by  two  or  four  rods  which  spanned  the  crank  shaft. 
The  cylinder  was  supported  by  A  frames.  The  pump  barrel 

was  bolted  to  the  base  of 
the  engine.  The  pump  was 
single  acting  and  the  general 
arrangement  was  excellent  and 
simple. 

This  type  o*  engine  was 
installed  in  1868  and  1876  at 
the  Eastbourne  Water  Works 
in  England.  It  proved  to  be 
so  successful  in  operation  and 
needed  so  little  repairing  after 
being  in  service  until  1881 
that  new  engines  for  that 
plant  were  made  of  the  same 
type. 

The  general  arrangement 
of  steam  end  and  pump  for 
this  machine  will  be  seen  to 
be  quite  similar  to  the  modern 
American  pumps.  The  design 
was  good  and  one  of  the 
simplest  for  the  application  of 
the  fly  wheel  to  the  pump. 

About  this  time  a  unique 
FIG.  65.— Fly  Wheel  Pump.  pump  was  introduced  by  Bird- 

sill  Holly  to  care  for  his  direct-pressure  system,  which  dispensed 
with  a  reservoir  or  standpipe,  obtaining  pressure  direct  from 
the  engine.  He  introduced  the  system  in  1866  in  Lockport, 
N.  Y.,  using  a  pump  driven  by  a  water  wheel,  but  in  1871  he 
used  his  quadruplex  engine  at  Dunkirk,  N.  Y.  The  four  pumps, 
Fig.  66,  were  placed  in  tandem  with  the  steam  cylinders,  which 
were  arranged  in  pairs,  each  acting  on  a  crank.  The  two 


RECENT  HISTORY 


69 


I 


70 


PUMPING  MACHINERY 


cranks  were  arranged  at  135°  and  the  frame  of  the  engine  was 
such  that  the  center  lines  of  the  engines  intersected  at  90°. 
The  figure  shows  the  construction  of  the  engine. 

Such  an  engine  was  necessary  for  this  system,  as  the  engine 
had  to  start  from  any  position  as  soon  as  the  drop,  in  pressure 
in  the  mains  moved  the  governor.  This  system  saved  the 
expense  of  a  reservoir,  but  the  necessity  of  keeping  the  engine 
under  steam  continually  made  the  steam  use  and  the  labor 


FIG.  67. — Leavitt's  Lynn  Pump. 

expense  very  high.  The  engine  could  be  run  with  any  number 
of  steam  or  water  cylinders,  and  the  engines  could  be  used  as 
single-expansion  cylinders  or  by  exhausting  from  one  to  the 
other  three,  as  was  done  in  1874  at  Rochester,  the  advantages 
of  compounding  could  be  had.  On  account  of  the  intermittent 
action  of  these  engines  their  duty  was  not  better  than  that  of 
other  pumps.  A  duty  of  86,176,315  foot-pounds  was  obtained 
on  the  6,000,000  gallon  pump  at  Buffalo,  N.  Y.,  in  1879. 

The  first  important  compound  steam  end  for  water  works 
in  America  was  designed  in  1872  by  Frederick  Graff,  chief 


RECENT  HISTORY  7 1 

engineer  of  the  water  department  of  the  city  of  Philadelphia. 
Compound  pumps  were  built  by  Simpson  &  Co.  of  England, 
in  1848,  although  most  of  the  pumping  engines  were  of  single- 
cylinder  type. 

This  Graff  engine  was  followed  in  1873  by  an  engine  designed 
by  Mr.  E.  D.  Leavitt,  jr.,  which  was  quite  similar  to  the  Graff  en- 
gine in  its  theoretical  operation,  although  different  in  its  arrange- 
ment. It  was  intended  for  the  water  works  of  Lynn,  Mass. 

As  shown  in  Fig.  67,  this  had  two  inclined  steam  cylinders 
and  a  pump  which  was  arranged  to  discharge  on  each  stroke, 
although  it  was  single  acting  on  the  suction  side.  This  was 
accomplished  by  the  plunger  attached  above  the  pump  bucket. 
The  supplementary  pipe  and  valves  connecting  the  two  ends 
of  the  pump  cylinder  were  for  the  purpose  of  reducing  the 
friction.  The  valves  were  of  large  diameter  and  double  beat, 
that  is,  each  valve  had  two  discharging  edges.  This  type  of 
pump  end  was  first  built  in  1848  and  was  known  as  a  Thames- 
Ditton  pump.  The  figure  also  shows  the  application  of  the 
fly  wheel  to  the  pump  for  the  purpose  of  using  steam  expan- 
sively. The  cylinders  were  steam  jacketed  and  the  steam 
valves  were  of  the  gridiron  type,  driven  by  cams.  The  double- 
acting  air  pump  was  driven  from  the  beam. 

This  pump  gave  a  duty  of  almost  104,000,000  foot-pounds 
per  100  pounds  of  coal  on  a  52-hour  test,  when  pumping  about 
5,060,000  gallons  per  day.  The  duty  for  its  year's  record  was 
about  75,000,000.  The  dimensions  of  the  pump  are  given  below: 

Diameter  of  high-pressure  cylinder 17 1  ins. 

low-pressure  cylinder 36    ' 

high-pressure  piston  rods 3    ' 

low-pressure  piston  rods 3 1  ' ' 

air  pump *. i  T  \ •" 

pump  barrel 26.1  ins. 

plunger « 18^ 

Length  of  stroke  of  steam  and  water  pistons 7  ft. 

air  pump 44^  ins. 

Distances  between  end  centers  of  the  beam 1 1  ft. 

Weight  of  fly  wheel 10.7  tons 

beams 4.2 

moving  parts  connected  with  beams 5 

Length  from  pump  to  top  of  vertical  pipe  reservoir. J904  ft. 

Height  of  top  of  vertical  pipe  above  bottom  of  well 163 .34  ft. 


72 


PUMPING   MACHINERY 


This  pump  with  its  walking  beam  may  be  taken  as  an 
example  of  many  of  the  pumping  engines  to  be  found  in  Europe 
and  America,  although  the  arrangement  of  the  cylinders  in  an 
inclined  position  was  novel.  Practically  all  of  these  machines, 
with  the  exception  of  Graff's  for  the  Philadelphia  Water  Works, 


FIG.  68.— English  Pump  of  1866. 

were  with  single  cylinders,  although  compound  engines  had 
been  known.  To  give  an  idea  of  one  of  the  English  pumps  of 
this  period  the  West  Middlesex  pump  of  1866  is  shown  in  Fig. 
68,  as  this  illustrates  the  general  arrangement  of  such  pumps. 
The  steam  cylinder  was  80  X 120  inches,  while  the  water  cylinder 
was  24^  X 120  inches.  The  pump  was.double  acting,  although  the 
steam  cylinder  was  single  acting.  The  large  acorn-shaped  box 


kECRNT  HISTORY  73 

B  on  the  pump  rod  directly  beneath  the  parallel  motion  is  the 
balance  box,  which  was  loaded  so  as  to  force  the  water  from 
the  pump  against  a  head  of  200  feet  on  the  down  stroke.  During 
this  stroke  the  two  ends  of  the  steam  cylinder  were  connected 
through  a  15-inch  equilibrium  pipe  as  in  the  old  type  of  Watt 
engine.  The  weighted  box  pulled  the  piston  up  until  the  piston 
reached  the  top  of  the  zo-foot  stroke,  when  the  plug  rod  R 
reversed  the  valves,  bringing  live  steam  on.  top  of  the  piston 
and  connecting  the  lower  end  of  the  cylinder  to  the  condenser 
C  through  a  ig-inch  exhaust  valve.  The  steam  on  the  upper 
•side  then  drove  the  piston  downward,  the  air  pump  A  driven 
'from  the  main  beam  maintaining  the  vacuum  in  the  condenser. 
This  pump  was  provided  with  a  safety  device  so  that  pressure 
would  be  held  in  the  discharge  chamber  in  case  the  discharge 
pipe  should  break.  This  pump  is  the  equivalent  of  Watt's 
original  pump,  although  it  was  worked  with  steam  at  40  pounds 
pressure.  The  water  pressure  was  200  feet  and  the  pump  made 
1 6^  strokes  a  minute. 

Several  important  inventions  of  Worthington's  should  be 
mentioned  here:  the  dash  relief  valve,  the  rotary  pump  valve, 
and  the  cross  connection  for  compound  pumps. 

The  dash  relief  valve  was  used  on  pumps  to  regulate  the 
length  of  the  stroke  and  prevent  pounding.  The  motion  of  the 
piston  could  be  stopped  by  cutting  off  the  exhaust  steam  before 
the  end  of  the  stroke  and  compressing  the  steam  in  the  space 
behind  the  pistons.  This  was  done  by  having  the  piston  ride 
beyond  the  port  to  the  exhaust  passage,  but  in  order  to  intro- 
duce live  steam  into  the  cylinder  in  this  cushion  space  it  was 
nedessary  to  use  another  passage  beyond  the  one  for  the  exhaust. 
This  made  five  steam  passages  in  the  cylinder,  Fig.  69,  the  two 
outer  for  steam  and  the  three  inner  ones  for  exhaust.  It  is 
seen  from  the  figure  that  when  the  piston  reaches  the  position 
shown,  the  steam  to  the  left  of  the  piston,  retained  behind  the 
piston  after  passing  B,  has  been  compressed  into  the  clearance 
space  and  into  the  passage  A.  This  acts  as  a  cushion,  and  as 
soon  as  the  valve  moves  to  the  right  high-pressure  steam 
enters  and  drives  this  piston  to  the  right.  If  the  piston  is 


74 


PUMPING  MACHINERY 


brought  to  rest  too  far  from  the  end  of  its  stroke,  a  valve  D 
is  opened  slightly  between  the  passages  A  and  B  and  some  of 
the  steam  compressed  in  A  exhausts  into  B  and  then  to  C, 


FIG.  69. — Dash  Relief  Valve. 

thus  allowing  the  piston  to  travel  farther.  If  D  is  opened  too 
much  the  piston  will  strike  the  cylinder  head,  causing  pounding, 
and  it  is  then  necessary  to  close  the  valve  D  slightly.  The 


FIG.  70. — Rotary  Steam  Valve. 


FIG.  71. — Cross  Connections. 


application  of  the  rotating  valve  was  the  application  of  the 
common  valve  used  by  Corliss  to  the  steam  cylinder  of  the  pump, 
Fig.  70.  This  valve  operates  in  the  same  manner  as  a  slide 


RECENT  HISTORY 


75 


valve,  the  steam  or  exhaust  being  conducted  by  the  passage 
A  or  the  center  of  the  valv6  casting  B.  The  dotted  passage 
leading  to  the  end  controlled  by  a  dash  relief  valve  acts  as  the 


.76 


P  UMPI NG  M  A  CHI  NER  Y 


1 

J 


passage  for  the  starting  steam.1  In  such  an  arrangement  the 
piston  starts  slowly  because  all  of  the  steam  to  start  it  would 
have  to  pass  the  relief  valve. 


kLCENT  HISTORY  77 

The  cross  connection  of  the  low-pressure  cylinders,  Fig.  71, 
is  one  which  is  used  to  supply  steam  for  the  demands  of  eithei 
cylinders  C  or  D  from  the  exhaust  of  both  of  the  high-pressure 
cylinders  A  or  B.  Without  the  cross  connection,  the  steam 
from  A  would  be  used  only  in  C  and  that  of  B  in  D,  but  by 
this  cross  connection  of  the  valve  chests  a  steady  motion  is 
obtained. 

The  duplex  Worthington  pump  was  also  made  in  the  com- 
pound form,  and  one  tested  in  1873  at  the  Belmont  Water  Works 
of  Philadelphia  gave  a  duty  of  54,416,694  foot-pounds  on  100 
pounds  of  coal.  This  pump  had  steam  cylinders  of  29  and 
50^  inches  in  diameter,  while  the  water  cylinder  was  22^-  inches 
in  diameter.  The  common  stroke  was  50  inches.  The  capacity 
of  the  pump  was  5,000,000.  Fig.  72  shows  the  general  form 
of  this  pump.  A  similar  one  at  Newark,  N.  J.,  gave  77,358,478 
foot-pounds  duty. 

The  difference  between  the  duty  of  the  Leavitt  pumping 
engine  and  that  of  the  direct-acting  duplex  Worthington  is 
rather  marked,  arid  where  the  engine  is  to  run  at  full  capacity 
for  considerable  length  of  time  these  figures  show  conclusively 
that  such  a  high-duty  engine,  even  though  its  cost  is  much 
greater  than  the  simpler  pump,  would  prove  to  be  a  paying 
investment.  Howrever  this  may  be,  when  the  use  of  the 
pump  is  quite  intermittent,  if  the  pump  is  too  large  for  the 
average  consumption  of  water,  the  loss  in  starting  up,  together 
with  interest,  depreciation,  insurance,  taxes,  and  repairs 
on  the  more  expensive  machine  might  make  the  cost  of 
pumping  greater  than  it  would  be  with  the  less  efficient  ma- 
chine. 

This  important  fact  was  emphasized  by  Mr.  Worthington 
and  undoubtedly  accounts  for  the  greater  use  of  these  duplex 
machines  in  the  years  which  immediately  followed. 

At  the  Centennial  Exhibition  of  1876  the  Worthington 
pump  was  used  to  supply  all  of  the  water.  The  pump 
is  shown  in  Figs.  72  and  73.  The  dimensions  were  as  fol- 
lows: 


78  PUMPING  MACHINERY 

Diameter  high-pressure  cylinder 29    ins. 

low-pressure  cylinder 50} 

water  plunger 22^ 

Stroke 48 

Air-pump  diameter 29! 

stroke 24 

From  the  sectional  view  of  the  pump  it  is  seen  that  the 
steam  and  water  cylinders  are  in  tandem  on  each  side  of  the 
pump  and  that  the  valves  are  of  the  swinging-piston  balanced 
type.  The  cylinders  have  separate  steam  and  exhaust  pas- 
sages and  the  outside  cut  shows  the  adjusting  relief  valves. 
The  cylinders  were  jacketed  on  the  barrel  and  heads.  It  will 
be  noted  that  the  valves  are  driven  from  bell-crank  levers 
attached  to  the  walking  beams  of  the  air  pumps  by  connecting 
rods.  The  air-pump  beams  are  worked  from  cross  heads  on  the 
piston  rods  through  short  connecting  rods.  It  is  evident  that 
the  connections  of  the  rods,  beams,  and  levers  are  such  that  the 
motion  of  one  engine  controls  the  valve  of  the  other  engine. 

The  construction  of  the  jet  condenser  and  the  air  pump  is 
seen  in  the  picture.  The  water  end  of  the  pump  shows  the 
small  valves  advocated  by  Worthing  ton  as  well  as  the  use  of 
the  plunger  and  ring.  The  direct  path,  of  the  water  through 
the  pump  is  one  which  diminishes  the  friction  loss  in  the 
machine. 

The  centennial  year  1876  marks  the  installation  of  another 
Leavitt  pump  at  Lawrence,  Mass.  This  pump  was  built  on  the 
same  lines  as  the  Lynn  pump,  but  it  gave  a  duty  of  117,550,800 
foot-pounds  per  100  pounds  of  coal,  making  a  new  record. 

The  Corliss  engine  of  the  Centennial  Exposition  was  one  of 
the  most  remarkable  features  of  the  exhibition.  In  1878  Mr. 
George  H.  Corliss  used  hir  engine  in  the  construction  of  a  pump 
for  Pawtucket,  R.  I.  This  pump,  Fig.  74,  consisted  of  two 
steam  cylinders,  each  one  in  tandem  with  a  water  cylinder. 
The  tail  rod  from  the  water  end  was  attached  by  a  connecting 
rod  to  a  vibrating  lever,  pivoted  in  the  base  and  joined  by  a 
rod  to  the  crank  of  the  engine.  By  use  of  the  fly  wheel  in  this 
and  the  other  fly-wheel  engines,  steam  could  be  used  expansively 
without  the  use  of  heavy  reciprocating  parts.  The  fly  wheel 


RECENT  HISTORY 


79 


was  mounted  on  the  shaft  carried  by  the  bearings,  which  were 
supported  by  the  air  chambers  of  the  pump.  A  diagonal 
brace  was  carried  from  the  bearings  to  the  main  center  ped- 
estals. 

The  steam  cylinders  were  15  and  30 J  inches  in  diameter, 
30  inches  in  stroke,  and  were  furnished  with  the  Corliss  valve 
gear.  They  were  steam  jacketed  on  the  barrels,  heads,  and 


FIG.  74. — Corliss  Pump  for  Pawtucket. 

valve  boxes.  There  was  a  receiver  between  the  cylinders,  the 
volume  of  which  was  equal  to  that  of  the  low-pressure  cylinder. 
The  drips  from  the  jackets  were  delivered  into  the  boiler  feed, 
heating  it,  while  the  drip  from  the  receiver  was  passed  through  a 
coil  in  the  boiler  flue  and  returned  to  the  receiver  in  a  super- 
heated condition.  The  steam  throttle  valves  were  so  connected 
with  the  governor  that  they  limited  the  speed  to  52  revolu- 
tions. 


80 


PUMPING   MACHINERY 


FIG.  75. — Moreland's  Compound  Steam  End. 

The  exhaust  from  the  low-pressure  engine  was  carried  to  a 
jet  condenser  placed  beside  the  engine,  while  the  air  pump  was 


RECENT  HISTORY  81 

placed  below  the  main  pedestal  bearing.  The  pump  was 
10.52  inches  in  diameter  with  2|-inch  rods.  It  was  to  lift 
3,000,000  gallons  per  twenty-four  hours  against  270  feet  head. 
It  was  of  the  inside-packed  plunger  type  with  280  small 
valves. 

In  August,  1878,  this  engine  was  tested  for  ten  hours  per 
day  for  twelve  days,  with  all  coal  and  wood  charged,  and  gave 
a  duty  per  100  pounds  of  coal  of  104,357,654  when  delivering 
3,060,000  gallons.  On  October  3,  1878,  a  24-hour  test  gave 
133,522,060  foot-pounds.  This  was  a  remarkable  result,  but 
no  more  so  than  the  annual  duty  of  123,656,000  foot-pounds 
for  the  year  1888. 

In  1889  Professor  Denton  tested  this  engine  and  obtained 
a  duty  of  124,720,000  foot-pounds  with  hard  coal  and  127,- 
350,000  foot-pounds  with  soft  coal.  These  results  were  obtained 
with  boilers  giving  an  equivalent  evaporation  of  8.88  pounds 
of  water  per  pound  of  coal  for  the  hard  coal  and  9.35 
pounds  for  the  soft  coal.  A  higher  duty  could  be  credited 
to  the  engine  if  the  usual  amount  of  10  pounds  were  evap- 
orated. 

The  next  important  water-works  engine  was  that  used  at 
the  Pettaconsett  Water  Works  of  Providence,  R.  I.  This  was 
built  by  Corliss  originally  for  Boston,  but  was  not  accepted  for 
some  reason.  Its  test  for  six  days  in  May,  1882,  gave  a  duty 
of  113,271,000  foot-pounds. 

The  inverted  pump  of  Moreland  and  Thompson  of  1868 
proved  to  be  so  successful  that  in  1880  Moreland  &  Son 
designed  a  compound  pump  for  the  Eastbourne  Water  Works 
of  England.  This  pump,  Fig.  75,  was  of  the  close  tandem  com- 
pound steam  type  with  a  single-acting  piston  pump  attached 
below.  The  plunger  on  the  piston  rod  as  shown  in  the  earlier 
pump,  Fig.  65,  was  arranged  to  give  a  discharge  on  each  stroke. 
The  general  arrangement  may  be  seen  in  the  figure  and  reminds 
one  of  the  lines  of  recent  pumps.  The  valve  gearing  was  of  the 
Myer  type  and  the  high-  and  low-pressure  cut  off  could  be 
adjusted  separately. 

The  general  dimensions  were  as  follows: 


82  PUMPING   MACHINERY 

High-pressure  diameter 20    ins. 

Low-pressure  diameter 38^    ' 

Pump-piston  diameter 20 

Plunger  diameter 15      " 

Stroke 40 

Fly-wheel,  15  tons 15  ft.  7  ins.  diam. 

This  pump  was  tested  on  February  18  and  19,  1884,  by 
Messrs.  Wallis  and  Borias,  and  showed  a  duty  of  124,600,000 
foot-pounds  per  112  pounds  of  coal  or  111,300,000  per  100 
pounds  of  coal.  It  was  running  under  a  low  head  during  this 
test  arid  developed  only  90.7  H.P.,  while  it  was  designed  to 
develop  160  H.P.  The  following  data  are  given  from  the 
test: 

Total  revolutions  in  twelve  hours 16,875 

Average  revolution  per  minute 23.44 

Total  amount  of  water,  in  gallons 759,456 

Average  total  lift,  in  feet 243.94 

steam  pressure  above  atmosphere 69.15 

vacuum 28.58  iris. 

Horse-power  water 77-97 

Indicated  horse  power. ." 90 .  70 

Mechanical  efficiency 86% 

Coal  consumed  in  twelve  hours 1666  Ibs. 

Coal  per  horse-power  hour  of  water i .  78 

Coal  per  indicated  horse-power  hour i .  53 

Duty  per  112  Ibs.  of  coal  in  ft.  Ibs 124,600,000 

Duty  per  100  Ibs.  of  coal  in  ft.  Ibs 111,300,000 

At  this  time  the  Holly  Manufacturing  Company  brought  out 
a  high-duty  engine  to  replace  the  quadruplex  engine.  This 
was  designed  by  Mr.  Harvey  F.  Gaskell,  their  superintendent. 
It  was  a  simpler  machine,  and  was  the  first  standard  design  of 
high  duty  applicable  to  all  forms  of  water  works,  the  previous 
high-duty  engines  being  of  special  design. 

The  first  of  these  engines  (Figs.  76,  77,  and  78)  was  installed 
at  Saratoga  Springs,  N.  Y.,  in  1882. 

From  the  figures  it  will  be  seen  that  the  high-pressure 
cylinder  is  directly  over  the  low-pressure  cylinder  and  that 
the  piston  rods  of  these  two  cylinders  are  connected  to  the 
extreme  ends  of  a  massive  walking  beam  by  short  connecting 
rods.  This  connection  means  that  as  one  piston  moves  to 
the  right  the  other  moves  to  the  left,  each  being  at  the  dead 


RECENT  HISTORY 


83 


i 


point  at  the  same  instant,  and  therefore  it  was  of  the  Wolf 
compound  type.     This  arrangement  made  it  possible  to  dis- 


PUMPING  MACHINERY 


charge  directly  from  the  high-pressure  cylinder  to  the  low- 
pressure  through  the  gridiron  valve  seen  in  Fig.  78. 


RECENT   HISTORY 


85 


A  connecting  rod  from  the  upper  end  of  the  beam  joined 
it  to  the  crank  of  the  fly-wheel  shaft  and  by  connecting  the 
two  sides  or  engines  by  cranks  at  right  angles  the  motion  of 


OUTSIDE.  SECTIONAL. 

-STEAM     END    ELEVATION,    GASKILL    PUMPING    ENGINE. 

FIG.  78. — Section  of  Gaskell  Engine. 

the  pumps  was  made  steady.  The  valve  gear  is  driven  by 
two  longitudinal  revolving  shafts  driven  by  bevel  wheels,  the 
high-pressure  steam  valve  being  of  the  poppet  type  with  an 


86  PUMPING  MACHINERY 

automatic  relief  while  the  other  valves  are  of  the  gridiron 
type.  The  condenser  can  be  driven  from  a  lever  attached  to 
the  trunnions  of  the  walking  beam,  although  in  the  Saratoga 
engine  a  Buckley  condenser  was  used.  The  water  end  con- 
sisted of  an  inside  packed  plunger,  the  Worthington  system 
of  a  large  number  of  small  valves  being  used  in  it.  In  this 
engine  there  were  672  water  valves.  There  were  two  tests 
made  on  this  pumping  engine;  one  in  November,  1882,  and 
the  other  hi  June,  1883;  both  of  the  results  showed  a  duty  of 
about  113,000,000  foot-pounds,  although  for  a  short  period  a 
duty  of  127,000,000  was  obtained.  The  dimensions  of  this 
engine  were: 

Diameter  high-pressure  cylinder 21  ins. 

low-pressure  cylinder 42    ' 

pump  plunger 20    " 

rods 4    " 

St-x5ke 36    " 

Capacity • 5,000,000  gals. 

There  were  subsequently  many  changes  made  in  the 
detaik  of  this  engine,  but  the  general  plan  was  unaltered. 
Charles  T.  Porter  in  his  report  of  1883  on  the  Saratoga  engine 
makes  the  following  statement:  "  In  the  details  through 
which  this  general  plan  has  been  carried  out,  there  are  no 
features  which  seem  open  to  criticism,  but,  on  the  contrary, 
all  seem  entitled  to  commendation.  The  construction  is 
thoroughly  mechanical  in  every  respect.  The  forces  are  trans- 
mitted and  the  strains  are  resisted  in  the  manner  theoretically 
the  most  correct.  ...  I  do  not  think  that  the  study  of  any 
machine  has  ever  given  me  a  stronger  feeling  of  confidence  in 
its  durability."  This  shows  the  kind  of  designing  done  by 
Mr.  Gaskell  and  the  kind  of  work  done  by  the  Holly  com- 
pany. 

While  these  pumps  were  being  built  in  America,  the  Euro- 
pean practice  seemed  to  hold  to  the  Cornish  steam  pump 
(Fig.  79),  although  in  many  cases  horizontal  fly-wheel  engines 
were  connected  through  other  rods  or  gears  to  horizontal  or 
vertical  pump  cylinders  (Figs.  80  and  81),  and  in  certain  cases 
the  beam  engine  was  built  with  a  fly  wheel  (Figs.  82  and  83). 


RECENT  HISTORY 


87 


FIG.  79. — Cornish  Engine  of  1878. 


FIG.  80. — Horizontal  Fly  Wheel  Pumping  Engine. 


88 


PUMPING  MACHINERY 


The  pump  shown  in  Fig.  83  was  one  constructed  by  Simpson 
&  Co.,  for  the  Lambeth  Water  Works  of  London.  This  pump 
was  of  the  type  used  for  many  years  by  this  company  and  gave 
unusually  good  results.  The  steam  end  consists  of  a  com- 
pound engine  connected  to  the  beam  by  a  Watt  parallel 
motion.  A  fly  wheel  was  used  to  replace  the  heavy  bob 
weight. 

The  technical  press  of  that  time  does  not  record  any  remark- 


FIG.  81.— 1874  Fly  Wheel  Pump. 

able  duties  for  these  machines,  and  for  that  reason  these  are  not 
described  here,  although  the  student  who  cares  for  examples  is 
referred  to  the  bibliography  at  the  end  of  the  book. 

An  exception  to  the  statement  above  must  be  mentioned— 
the  pump  used  for  the  city  of  Paris  at  the  St.  Maur  station. 
Although  installed  in  1876,  it  was  somewhat  similar  to  one 
installed  in  1871.  They  were  built  by  Farcot  &  Sons.  The 
steam  cylinder  was  39.4X70.8  inches,  and  the  water  cylinder 
was  14,2X70.8  inches.  The  distance  from  the  center  of 
the  shaft  to  the  center  of  the  water  cylinder  was  50  feet 
8  inches.  The  pump  had  a  piston  speed  of  350  feet  per 


RECENT  HISTORY 


89 


90  PUMPING  MACHINERY 

minute,  which  meant  about  30  R.P.M.,  a  high  speed  for  that 
day.  At  this  speed  it  pumped  3,000,000  gallons  in  twenty- 
four  hours.  To  get  such  a  high  speed  the  valves  on  the  suc- 


FIG.  83. — Simpson's  Pumping  Engine. 

tion  and  discharge  side  were  made  so  as  to  open  simply  and 
give  a  free  passage  to  the  water.  This  may  be  seen  in  the 
figure.  The  passages  through  the  pump  were  made  quite  large 
so  that  the  water  would  have  a  free  path. 


RECENT  HISTORY 


91 


92 


PUMPING  MACHINERY 


The  extended  tail  rod  on  the  pump  plunger  was  used  to 
pump  air  into  the  air  chambers  to  keep  them  charged. 

The  pump  was  driven  by  a  single-cylinder  condensing 
engine,  with  the  air  pump  driven  by  a  lever  connected  to  the 


FIG.  85. 

cross-head,  and  so  well  was  the  machine  designed  that  the 
makers  reported  the  development  of  an  indicated  horse  power 
on  12.  i  pounds  of  steam  per  hour  or  1.54  pounds  of  ordi- 
nary coal  while  2.03  pounds  _of  coal  were  used  per  pump  horse 


RECENT   HISTORY 


93 


power  per  hour.     This  would  mean  a  duty  of  78,000,000  foot- 
pounds per  100  pounds  of  coal. 

The  6,ooo,ooo-gallon  Reynolds  pump  of  1881  (Figs.  85  and 
86)  for  Milwaukee  shows  a  type  similar  to  many  of  the  European 
pumps  of  this  period  in  its  general  arrangement  of  cylinders, 
except  that  there  was  an  additional  low-pressure  cylinder 


FIG.  86. — Reynolds  Milwaukee  Engine. 

added  above  the  beam.  The  duty  of  this  engine  was  104,000,000 
foot-pounds  per  100  pounds  coal.  Following  this  came  the 
two  12,000,000  gallon  Allegheny  pumps  of  1883  (Fig.  87), 
in  which  there  were  three  cylinders  to  each  pump — one  high 
pressure  and  two  low  pressure,  The  engine  was  an  entirely 


94 


PUMPING  MACHINERY 


new  design  for  large  engines,  although  small  pumps  had  been 
arranged  in  this  general  manner  much  earlier.  The  cranks 
were  placed  at  120°  and  the  number  of  working  steam  cylin- 


FIG.  87. — Allegheny  Pump. 

ders  was  reduced  to  three  in  place  of  four,  so  common  in  the 
standard  type  of  pump  of  the  day.  The  arrangement  of  the 
pump  and  steam  cylinders  is  clear  from  the  figure.  The  valve 


FIG.  89. — Richardson' 


pie-expansion  Engine. 

(To  face  page  94) 


RECENT  HISTORY  95 

gear  was   of  the  Corliss  type  with  fly  ball  and  hand  govern- 
ing   and    receivers    were    used    between    the   cylinders.      The 


FIG.  88. — First  Triple  Expansion  Pump. 

cylinders  were  jacketed,  tmt  there  was  no  heating  coil  in  the 
receiver. 

The  valves  of  the  pumps  were  originally  of  the  Cornish 


96  PUMPING  MACHINERY 

double-beat  design,  but  these  were  later  changed  to  small 
valves  supported  on  a  cage  or  basket  placed  over  the  opening 
for  the  Cornish  valves. 

The  test  of  this  engine  gave  107,000,000  foot-pounds  per 
loop  pounds  of  steam  when  tested  in  1884  by  Mr.  C.  A.  Hague 
and  Professor  David  M.  Green.  It  represented  a  type  of 
engine  of  great  value  for  small  floor  space,  as  was  the  case  of 
the  Eastbourne  engine  of  1880. 

After  the  Allegheny  engine  of  1883  Allis  built  several 
compound-beam  fly-wheel  engines  of  about  2,500,000  gallons, 
and  in  one  of  them  for  Hannibal,  Mo.,  the  clearance  was 
reduced  by  putting  the  Corliss  valves  in  the  head  of  the 
cylinder.  This  engine  on  test  in  November,  1885,  by 
Mr.  C.  A.  Hague,  gave  a  duty  of  118,327,025  foot- 
pounds per  100  pounds  coal  at  a  gauge  pressure  of  79 
pounds.  This  engine  then  led  to  the  design  of  the  triple- 
expansion  engine. 

The  triple-expansion  engine  (Fig.  88)  is  the  first  of  the  most 
popular  type  of  large  pumping  engine,  although  many  other 
forms  have  been  suggested  and  used.  This  pump  was  built 
for  the  city  of  Milwaukee.  The  figure  shows  clearly  the 
arrangement  of  the  pump  and  engine.  This  resembles,  to  some 
extent,  the  design  of  frame  used  by  Moreland  in  England. 
It  is  in  reality  the  same  arrangement  as  that  used  for  'marine 
engines  with  pumps  added  below  the  bed  plate  of  the 
engine. 

The  engine  gave  the  high  duty  of  122,483,204  foot-pounds 
per  100  pounds  of  coal  with  80  pounds  steam  pressure.  The 
dimensions  of  the  engine  were  as  follows: 

High-pressure  cylinder  diameter 21    ins. 

Intermediate- pressure  cylinder  diameter 36 

Low-pressure  cylinder  diameter 51 

Pump  plunger 23  \ 

Stroke 36 

Capacity  per  twenty-four  hours 6,000,000  gals. 

The  first  triple-expansion  pump' in  England  (Fig.  89)  was 
introduced  in  1891  by  S.  Richardson  &  Sons,  and  gave  a  steam 
consumption  of  13.53  pounds  per  I.H.P.  hour.  The  use  of 


RECENT  HISTORY 


97 


triple-expansion  for  direct-acting  pumps  was  also  introduced 
at  this  time  by  Davison  in  his  water- works  pumps.  These 
gave  good  results  and  increased  the  duty  of  direct-acting 
pumps. 

In  1885  Worthington   brought    out  his  high-duty  engine, 


built  exactly  as  the -duplex  engines,  but  with  the  addition  of 
compensating  cylinders  so  as  to  use  steam  expansively.  It 
was  not  possible  to  obtain  the  high  duties  of  the  expansive 
fly-wheel  pumps  with  the  ordinary  duplex  pump,  and  although 
these  could  be  sold  much  cheaper,  their  duty  was  so  low  as  to 


98 


PUMPING    MACHINERY 


prevent  their  use  in  many  cases.  The  use  of  heavy  recipro- 
cating weights  was  not  feasible,  and  Worthington  improved 
the  invention  of  Mr.  J.  D.  Davies  of  1879  by  which  steam  could 
be  used  expansively.  The  arrangement  as  applied  by  Mr. 
C.  C.  Worthington  is  shown  in  Fig.  90,  where,  as  the  piston 
moves  to  the  right,  the  tail  rod  of  the  pump  forces  the  plungers 
into  the  oscillating  cylinders  A  A  against  water  pressure  from 
the  cylinder  Z?,  to  which  they  are  connected  by  pipes.  The 
pipes  are  connected  through  stuffing  boxes  to  the  trunnions 


O  O  0^0  O  0 


FIG.  91. — Independent  Air  Tank  Compensator. 

of  the  cylinders.  It  will  be  seen  that  the  side  thrust  from  these 
is  balanced  on  account  of  their  symmetrical  position.  The 
cylinder  B  is  connected  to  the  top  of  the  air  chamber  in  the 
force  main  of  the  pump,  and  thus  the  pressure  of  the  air  cushion 
above  the  fluid  in  B  is  always  equal  to  that  in  the  force  main. 
By  the  arrangement  of  multiplying  cylinders  at  the  base  of 
B  the  unit  pressure  on  the  plungers  of  A  A  is  much  greater 
than  that  of  the  force  main,  as  the  area  of  the  plunger  is  much 
smaller  than  the  piston  of  B. 

The  pistons  are  opposed  by  the  pressure  in  the  compensators 


RECENT  HISTORY 


99 


until  the  engine  reaches  its  mid-position;  the  plungers  are  here 
in  a  vertical  position.  From  this  point  to  the  end  of  the  stroke 
the  pressure  on  the  plungers  is  aiding  the  motion,  the  exact 
amount  being  proportional  to  the  cosine  of  the  angle  of  incli- 
nation to  the  horizontal.  This  then  means  that  at  the  beginning 
of  the  stroke  the  steam  pressure  will  have  to  exceed  the  pressure 
on  the  main  pump  plunger,  diminishing  until  the  mid-stroke, 
and  from  there  on  it  may  be  less  than  the  pressure  on  the  main 
plunger.  This  permits  the  use  of  steam  expansively.  The 
application  of  an  independent  air  tank  (Fig.  91)  is  used  at 


FIG.  92. — Action  of  Compensator. 

times  with  this  system.     The  figure  gives  a  clear  idea  of  the 
arrangement  of  the  rams  and  plungers. 

To  better  illustrate  the  action  of  this,  Fig.  92  is  given;  in 
it  the  indicator  cards,  the  pump  cards,  the  various  positions  of 
compensating  cylinders  and  the  combined  cards  are  shown. 
At  the  lower  right-hand  side  of  the  figure  the  total  steam- 
pressure  card  has  been  placed  over  the  total  water-pressure 
card  .and  a  dotted  line  has  been  drawn  showing  the  variation 
of  total  pressure  exerted  horizontally  by  the  compensators. 
It  is  evident  that  during  the  first  half  of  the  stroke  the  excess 


100 


PUMPING  MACHINERY 


of  total  steam  pressure  just  equals  the  positive  resistance  from 


(i. 

I 


the  compensators,  while  during  the  latter  part,  the  excess  of 


RECENT    HISTORY-  101 

water  pressure  is  made  up  of  the  negative  resistance  or  returned 
energy  from  the  compensators.  This  enables  the  pump  to 
obtain  all  of  the  advantages  of  the  expansive  use  of  steam  and 
tests  of  these  engines  have  shown  this. 

Fig.  90  shows  the  application  of  this  in  1885,  wherein  the 
forms  of  water  end  and  steam  end  are  obvious.  The  steam 
end  is  jacketed  throughout  and  has  separate  steam  and  exhaust 
passages  with  rotary  cut-off  valves  in  addition  to  the  balanced 
slide  valves.  This  type  was  soon  changed  so  far  as  the  valve 
was  concerned,  but  the  compensating  principle  was  practically 
the  same. 

A  test  of  the  first  engine  of  this  type  built  in  1885  for  the 
water  works  of  New  Bedford,  Mass.,  gave  a  test  duty  of 
79,238,160  foot-pounds  on  100  pounds  of  coal,  and  this  test 
was  followed  by  another  by  Mr.  John  G.  Mair,  a  representative 
of  Simpson  &  Co.,  the  English  builders  of  pumping  machinery. 
Mr.  Mair  found  that  the  engine  used  from  14.53  pounds  to  15.05 
pounds  of  steam  per  I.H.P.  hour. 

This  engine  was  of  the  compound  type,  as  shown  in  the 
figure,  and  was  developed  extensively,  being  used  in  a  vertical 
position  in  places  where  ground  area  was  limited,  as  in  an 
installation  made  at  Memphis,  Tenn.  This  same  engine  was 
tested  in  1891  and  gave  a  duty  of  117,325,000  foot-pounds  per 
1000  pounds  of  steam. 

This  compensated  form  of  pump  was  also  developed  into 
a  triple-expansion  steam  end.  One  of  the  first  built  by 
Worthington  in  1891,  for  Turtle  Creek,  Pa.,  was  of  the  vertical 
form  with  the  high-duty  attachment.  The  triple-expansion 
forms  are  shown  in  Figs.  93  and  94. 

Fig.  93  shows  the  pump  installed  at  Fall  River,  Mass.,  for 
the  city  pumping  station  in  1909.  It  has  three  steam  cylinders, 
21,  33,  and  60  inches  in  diameter,  respectively,  arranged  in 
tandem  with  the  24^-inch  water  cylinder.  The  common 
stroke  is  36  inches.  The  high-duty  attachment  is  placed 
between  the  high-pressure  steam  cylinder  and  the  water  cylinder. 

The  pump  takes  its  suction  from  an  open  conduit  and 
pumps  directly  into  the  city  mains  at  about  100  pounds  per 


102  ,/VJ/P/;;^  MACHINERY 

square  inch.     The  capacity  was  10,000,000  gallons  per  twenty- 


four  hours.     The  pump  is  of  the  outside -packed  plunger  type 


RECENT  HISTORY  103 

with  special  glands.  The  cylinders  are  jacketed  and  reheat er 
coils  are  used  between  the  cylinders.  The  auxiliary  pumps  for 
charging  the  air  chamber  of  the  compensator,  the  air  pump, 
and  jacket  pump  are  driven  from  the  main  pump.  The  stroke 
governor  regulates  the  stroke,  keeping  it  at  its  full  value  by 
means  of  a  change  of  the  pressure  in  the  compensator. 

The  official  test  of  this  pump  gave  a  duty  of  136,500,000 
foot-pounds  per  1000  pounds  of  dry  steam.  A  pump  similar 
in  principle  to  the  above  was  installed  at  Montreal  in  1908  and 
gave  a  duty  of  177,538,000  foot-pounds  per  1000  pounds  of  steam. 

Fig.  94  gives  a  good  idea  of  the  vertical  form  as  installed 
at  the  Central  Park  pumping  station  in  Chicago,  although 
the  figure  is  not  of  this  particular  pump.  These  pumps  are 
duplex  with  steam  cylinders,  21,  33,  and  60  inches  in  diameter, 
and  of  5o-inch  stroke,  while  the  double-acting  water  cylinders 
are  34J-inch  diameter  and  5o-inch  stroke.  The  steam  cylinders 
are  jacketed  on  heads  and  barrels  and  the  valve  gearing  is  of 
the  four- valve  rotative  type.  At  18  revolutions  per  minute 
this  pump  lifted  20,000,000  gallons  per  twenty-four  hours. 
The  duty  on  test  was  174,735,801  foot-pounds  per  1000  pounds 
of  superheated  steam  used. 

By  means  of  the  balancing  plungers  shown  in  Fig.  94,  the 
weight  of  the  moving  parts  is  supported  by  plungers  BB,  which 
force  water  into  a  tank  C  against  air  pressure.  The  pressure 
in  this  tank  is  regulated  by  the  pressure  in  the  discharge  main. 
When  this  pressure  falls,  due  to  a  break  in  the  line,  the  weight 
of  the  moving  parts  would  not  be  supported  by  the  water 
pressure  and  the  pump  would  soon  come  to  rest.  The  steam 
pressure  is  used  to  force  the  water  on  each  stroke.  At  points 
A  A  on  the  piston  rods  between  the  low-pressure  cylinder  and 
the  water  cylinder  the  digh-huty  attachment  is  applied  which 
permits  of  the  expansive  use  of  steam,  as  shown  previously. 

The  suction  pipe,  which  is  connected  to  a  large  cistern  in 
the  pump  room,  is  joined  to  the  surface  condenser  so  that  all 
of  the  water  passes  through  the  tubes.  In  this  way  the  con- 
densation is  cared  for  in  a  station  situated  in  the  center  of  the 
city  where  condensing  water  is  not  at  hand. 


104 


PUMPING  MACHINERY 


One  of  the  latest  types  of  compensating,  direct-acting 
pump,  permitting  the  expansive  use  of  steam  without  the  use 
of  a  fly  wheel,  is  the  invention  of  Mr.  Luigi  d'Auria.  The 
pump  (Fig.  95)  has  a  cylinder  A  between  the  steam  and  water 
cylinders,  the  ends  of  which  form  the  extremities  of  a  pipe 
circuit  which  in  some  cases  is  used  as  the  base  of  the  pump. 
The  cylinder  A  contains  a  piston,  and  as  this  moves  back  and 
forth  the  water  with  which  the  system  is  rilled  is  compelled 
to  travel  first  in  one  direction  and  then  in  the  other. 

At  the  beginning  of  the  stroke  the  inertia  of  the  water 


FIG.  95. — d'Auria  Pump. 

utilizes  the  excess  steam  pressure,  while  in  bringing  the  water 
to  rest  it  will  exert  a  force  to  supply  the  deficiency.  In  this 
way  steam  may  be  used  expansively  in  the  steam  cylinder  to 
drive  the  pump  against  a  uniform  water  pressure.  The  steam 
may  be  used  in  one  or  more  cylinders,  depending  on  the  purpose 
of  the  engine.  When  multiple  expansion  is  wanted  it  can  be 
applied  to  this  engine.  The  compensation  is  so  perfect  with 
this  machine  that  high  speed  may  be  employed  with  no  danger 
of  pounding. 

This  method  has   also   been    applied    to    triple-expansion 
pumps  where  high  duties  have  been  sought. 


RECENT  HISTORY 


105 


About  the  end  of  the  last  century  Mr.  Charles  L.  Heisler 
of  Erie,  Pa.,  introduced  a  new  type  of  compensated  engine 
which  he  described  in  1900.  His  design  employs  a  duplex 
arrangement  of  cylinders,  as  shown  in  Fig.  96,  which  is  built  for 
triple  expansion.  The  line  drawings  show  that  each  piston  rod 
is  connected  to  a  vibrating  radius  rod  A  by  a  short  connecting 
rod  B  and  that  the  radius  rods  are  joined  by  the  link  C.  Just 
before  the  left-hand  engine  reaches  the  end  of  its  upper  stroke, 
as  in  position  i,  the  movement  of  the  link  C  in  compression 
lifts  the  left  radius  rod  and  aids  in  the  movement  of  that  pump. 
Position  2  shows  that  the  left  pump  is  being  aided  by  com- 
pression in  the  rod  C,  while  the  other  positions  show  the  pumps 


Position 


Position  2  Position  3  Position  4 

FIG.  96. — Heisler  Pump. 


at  the  two  remaining  starting  points.  In  each  of  these  posi- 
tions the  excess  of  pressure  at  the  beginning  of  the  stroke  is 
carried  over  to  the  other  side  and  aids  the  steam  when  that 
pressure  is  less  than  the  resistance  of  the  water.  This  method 
gave  fair  results  and  the  builders  claim  duties  of  130,000,000.  to 
160,000,000  foot-pounds  per  1000  pounds  of  steam,  although 
the  author  has  no  records  of  any  tests  giving  such  high  duties. 
The  method  of  having  one  side  of  the  pump  aid  the  other 
by  means  of  a  linkage  was  not  new  at  this  time,  -for  in  1874 
Fielding  of  England  proposed  such  a  scheme  and  in  1887  Henry 
Davey  patented  a  less  complex  form.  The  Davey  method  of 
compensation  is  given  in  Fig.  97.  In  this  system  single-acting 
water  plungers  are  connected  with  double-acting  steam  cylinders. 


106 


PUMPING  MACHINERY 


When  the  plunger  in  A  is  moved  to  the  right  there  is  little 
resistance  to  its  motion,  so  that  the  excess  of  steam  pressure 
is  carried  through  the  rod  D  to  the  plate  C  and  from  it  to  the 
other  side  of  the  pump  by  the  rod  E.  At  the  beginning  of  the 
stroke,  when  little  of  the  pressure  from  the  side  A  is  needed, 
it  is  seen  that  the  rod  D  has  little  leverage  about  the  center 
of  C  while  E  has  a  long  leverage.  When,  however,  the  plunger 
of  B  has  been  driven  to  the  left  and  the  steam  on  that  side  has 
expanded  so  that  more  pressure  is  needed,  the  leverage  of  the 
rod  D  becomes  greater,  giving  more  pull  on  the  rod  E.  At  the 
end  of  the  stroke  the  low  pressure  on  the  side  A  is  pulling  with 
full  leverage  and  exerting  a  great  force  on  E,  which  pulls  on 


FIG.  97. — Davey  Compensator. 

the  plunger  of  B.  B  requires  this  additional  pressure,  as  the 
steam  on  that  side  has  expanded  to  a  low  pressure.  The 
valve  gear  is  such  that  these  events  can  take  place  as  described, 
the  two  pumps  reversing  at  the  same  time.  The  pumps  des- 
cribed by  Davey  in  London  Engineering  in  1877  and  the  Watt 
engine  described  in  the  same  magazine  in  1885,  are  mentioned 
as  similar  to  the  above  in  principle. 

From  the  introduction  of  the  triple-expansion  fly-wheel 
pump  of  Allis  in  1886  to  the  present  there  have  been  a  number 
of  improvements  in  the  arrangement  of  steam  valves  and 
reheaters,  pump  cylinders  and  valves,  condensers  and  other 
details.  With  the  gradual  increase  in  the  steam  pressure,  the 


RECENT  HISTORY  107 

duty,  when  measured  on  the  basis  of  1000  pounds  of  steam, 
has  increased  together  with  that  of  the  duty  on  the  basis  of 
1,000,000  B.T.U.  The  type  first  used  by  Allis  has  been  adopted 
by  other  large  engine  builders  and  this  design  in  the  hands  of 
the  Worthington,  Blake,  Holly,  Snow,  and  Southwark  companies 
has  proven  it  to  be  a  good  one.  Mr.  E.  D.  Leavitt,  Jr.,  has  also 
designed  special  types  during  these  years  which  have  given 
excellent  results,  as  will  be  seen  later.  The  duties  on  the  basis 
of  1000  pounds  of  steam  have  gone  from  122,452,729  foot-pounds 
in  1886  to  154,048,704  foot-pounds  in  1892,  to  179,454,250  in 
1900,  and  finally  to  181,048,605  foot-pounds  in  1906.  This 
method  of  stating  the  duty  has  the  objection  that  the  efficiency 
is  dependent  on  the  amount  of  heat  in  each  pound  of  steam 
supplied.  The  amount  of  heat  in  1000  pounds  of  dry  steam 
depends  on  the  pressure  of  the  steam  and  consequently  it  is 
perfectly  possible  to  have  an  engine  with  a  smaller  duty  per 
1000  pounds  of  steam  more  efficient  actually  than  one  with  a 
a  higher  looo-pound  duty.  The  method  of  expressing  the  duty 
as  the  number  of  foot-pounds  per  million  British  thermal  units 
supplied  is  far  better,  as  in  this  case  the  real  efficiency  is  being 
measured. 

The  engine  giving  the  highest  actual  thermal  efficiency, 
according  to  C.  A.  Hague,  is  the  Norberg  quadruple-expansion 
engine  of  1899,  which  gives  22.80  per  cent;  however,  the  thermal 
efficiency  based  on  useful  work  delivered  by  the  pumps,  obtained 
by  Professor  Thurston  in  1899,  was  20.9  per  cent.  This  result 
was  within  84  per  cent  of  the  theoretically  perfect  engine. 
This  pump  was  built  for  the  Pennsylvania  Water  Company, 
near  Pittsburg.  There  were  four  steam  cylinders  as  well  as  the 
pump  barrels  and  fly  wheel.  The  pump  was  built  on  peculiar 
lines  because  the  water  level  of  Allegheny  River  is  subject  to 
much  variation,  and  moreover  the  pump  was  to  be  placed  with 
two  other  similar  6,ooo,ooo-gallon  pumps  within  a  well  of  38 
feet  diameter  belonging  to  the  company.  The  dimensions 
of  this  engine  are  as  follows: 


108  PUMPING  MACHINERY 

Diameter  high-pressure  cylinder 19  J  ins. 

first  intermediate  cylinder 29^    ' ' 

second  intermediate  cylinder 49  £    " 

low-pressure  cylinder 5  7  £    ' 

plunger  (double-acting)  cylinder 14!    " 

Stroke 42 

Fly-wheel  diameter • 13  ft. 

Capacity c 6,000,000  gals. 

This  engine  gave  the  highest  duty  on  the  absolute  basis, 
although  it  does  not  give  so  high  a  value  per  1000  pounds  of 
dry  steam  as  those  which  have  been  built  later;  the  best 
accepted  duty  on  this  basis  being  about  180,000,000  foot- 
pounds. 

Figs.  98  and  99  illustrate  two  modern  forms.  The  first 
is  a  pump  installed  in  1909  at  Brockton,  Mass.  The  steam 
cylinders  are  19,  36,  and  54  inches  in  diameter,  and  the  stroke 
is  36  inches.  At  40  R.P.M.  the  36-inch  single-acting  plungers 
will  lift  6,000,000  gallons  per  twenty-four  hours  against  300 
feet  head.  The  duty  of  this  pump  was  169,982,000  foot-pounds 
per  1000  pounds  of  dry  steam  at  150  pounds  pressure. 

The  second  pump  illustrates  a  horizontal  cross-compound 
condensing  crank  and  fly-wheel  Snow  engine  for  Elyria,  Ohio. 
The  capacity  of  the  double-acting  plungers,  14^-  inches  in  diam: 
eter,  is  5,000,000  gallons  per  twenty-four  hours  at  36  R.P.M. 
The  steam  cylinders  are  22  and  48  inches  in  diameter  with  a 
36-inch  stroke.  The  steam  pressure  was  120  pounds  per  square 
inch  and  the  total  water  pressure  happened  to  be  the  same.  The 
duty  was  137,000,000  foot-pounds  per  1000  pounds  of  dry  steam. 

These  two  pumps  illustrate  clearly  the  form  of  modern 
high-duty  pump  for  water  works.  Fig.  98  illustrates  the 
vertical  self-supporting  pump  in  which  the  valve  boxes  and 
air  chambers  form  the  base  for  the  pump,  and  the  pump  barrel  is 
a  separate  casting.  The  steam  valve  gearing  is  of  the  Corliss  type. 
The  valves  are  placed  in  the  sides  of  the  high-pressure  cylinder, 
in  the  sides  and  heads  of  the  intermediate,  and  in  the  heads 
of  the  low.  The  side  straddling  rods  leading  from  the  cross- 
head  to  the  end  of  the  plunger  so  as  to  clear  the  shaft  and 
crank  are  shown.  The  governors,  valve  controls,  and  oiling 
devices  are  also  apparent.  Such  pumps  are  usually  placed  in 
basements  so  that  the  level  of  the  main  floor  of  the  station  is 
just  over  the  air  chamber. 


RECENT.  HISTORY 


109 


Fig.  99  is  the  Snow  type  of  Horizontal  pump  designed  to 
take  up  little  floor  space  and  yet  have  the  cylinders  for  the 


FIG.  98. — Holly  Triple  Expansion  Pump. 

water  and  steam  ends  in  such  a  position  that  they  may  be 
easily  examined  and  repaired.     The  method  of  clearing  the 


110  PUMPING  MACHINERY 

crank  and  shaft  by  two  rods  joining  the  cross-head  to  the 
plunger  rod  is  clearly  shown  a?  well  as  the  arrangement  of 
air  chambers,  pump  cylinders,  steam  receiver,  and  valves. 

In  this  period  of  the  important  developrr^ent  of  water-works 
pumping  engines  strides  were  being  made  with  the  centrifugal 
pump.  As  will  be  remembered,  it  was  about  the  year  1846 
that  Andrews  showed  that  curved  vanes  were  better  than  the 
straight  vanes  in  the  old  Massachusetts  pump  (Fig.  39).  This 
development  was  carried  out  by  Gwynne  of  England.  At  this 
time  Andrews  also  suggested  the  value  of  having  the  vanes  of 


FIG.  99. — Snow  Compound  Pump. 

the  runner  inclosed  between  two  discs  as  shown  in  Fig.  40. 
In  1851  Appold  showed  the  advantage  of  the  curved  vane  of 
Lloyd's  fans  experimentally. 

From  now  on  the  centrifugal  pump  was  used  extensively 
and  the  patent  gazettes  are  filled  with  the  record  of  improve- 
ments. The  great  field  for  this  pump  was  the  lifting  of  large 
quantities  of  water  through  small  distances,  where  there  was 
not  much  floor  space  available,  and  also  where  the  original 
cost  of  the  apparatus  was  of  some  importance.  The  firm  of 
John  &  Henry  Gwynne  in  England  was  the  best  known  of 
the  early  builders,  and  their  pumps  for  the  drainage  of  the  low- 


RECENT   HISTORY  111 

lands  of  Holland  and  Denmark  and  the  emptying  of  large 
dry  docks  were  very  successful. 

Not  only  did  the  Gwynnes  try  low  lifts  of  15  and  30  feet, 
but  in  1868  they  built  a  pump  with  a  runner  2  feet  in  diameter 
to  lift  water  against  a  total  of  18  feet  .suction  and  114  feet 
discharge.  The  pump  was  driven  at  a  speed  of  910  revolutions 
per  minute. 

In  1869  these  pumps  were  applied  to  condensers  of  steam 
engines.  For  this  service  they  are  admirably  adapted,  as  a 
large  body  of  water  has  to  be  lifted  a  small  distance  in  the  case 
of  stationary  engines,  while  for  marine  practice  the  only  work 
the  pump  does  is  to  overcome  the  resistance  due  to  the  moving 
of  the  water  and  the  friction  of  the  pipes  and  tubes. 

The  early  study  of  these  pumps  was  quite  meager  and 
their  design  was  more  or  less  empirical,  but  as  their  use  extended 
a  better  understanding  of  them  was  had.  Although  inexpensive 
for  the  quantity  of  water  handled  and  apparently  of  a  form  to 
give  good  results  the  efficiency  of  this  machine  was  low.  The 
earlier  efficiencies  of  50  per  cent  were  increased  to  65  and  70 
per  cent  in  1885.  Some  records  give  as  high  as  75  per  cent  for 
this  type  of  pump.  High  efficiencies  can  be  had  only  by  careful 
design  and  construction. 

The  application  of  this  pump  to  other  services  increased. 
In  1869  it  was  employed  for  dredging,  although  a  patent  of 
Louis  Schwartzkopff ,  granted  in  England  in  1856,  foreshadowed 
this  application  of  this  simple  machine,  while  for  draining 
wells  vertical  shafts  were  used  in  1884.  Before  this  time, 
however,  the  manner  of  balancing  end  thrust  by  bringing  in 
water  from  each  side  was  employed.  The  possibility  of  handling 
large  quantities  of  water  is  illustrated  by  a  pump  built  by 
R.  Moreland  &  Sons  for  a  dock  at  Malta,  where  it  was  required 
to  lift  30,000  gallons  of  water  per  minute  against  a  compara- 
tively low  head.  This  was  built  in  1887.  Another  pump  was 
built  a  little  earlier  by  the  Southwark  Foundry  and  Machine 
Co.,  designed  to  raise  40,000  gallons  per  minute. 

W.  O.  Weber  showed  by  experiment  that  the  efficiency  of  a 
pump  decreased  after  a  certain  head  was  reached,  and  so  when 


112 


PUMPING   MACHINERY 


greater  heads  were  to  be  overcome  by  centrifugal  pumps, 
it  was  suggested  by  some  (W.  A.  Booth  claims  the  honor  of 
this)  that  two  pumps  be  placed  in  a  series  (Fig.  100).  This 
then  led  to  the  multi-stage  pump,  where  several  pumps  were 
brought  together  in  one  casting.  The  water  discharged  from 
one  stage  passes  over  to  another  stage,  where  its  pressure  is 
increased.  Continuing  in  this  manner  the  pressure  against 


FIG.  100. — Two-Stage  Pump. 

which  the  pump  will  act  effectively  can  be  made  as  great 
as  desired.  The  Allis-Chalmers  Company  have  recently 
installed  a  five-stage  pump  which  lifts  3000  gallons  of 
water  per  minute  against  a  total  head  of  about  700  feet. 
Further  details  of  the  centrifugal  pump  will  be  considered 
in  a  later  chapter. 

The  rotary  pump  has  been  the  subject  of  many  patents 
during  this  later  period.     The  principles  involved,   however, 


RECENT  HISTORY 


113 


were  all  embodied  in  the  earlier  designs.  A  few  will  be  described 
to  give  some  idea  of  the  progress  made  in  this  form  of  pump. 

The  Behrens  pump  of  1867  (Fig.  101)  represents  a  form  of 
rotary  pump  using  rotating  lobes  with  no  sliding  parts.  The 
shafts  A  A  are  geared  to  move  in  opposite  directions  with  the 
same  velocity.  The  pistons  B  are  so  formed  that  they  slide 
over  the  bore  of  the  main  cylinder  C  and  the  inner  abutment 
liners  Z),  which  are  stationary.  These  liners  are  carried  on  the 
heads  on  one  side  of  the  cylinder,  while  the"  pistons  B  are  carried 
from  the  shaft  A  by  means  of  a  flange  cast  with  the  pistons. 
This  is  seen  better  in  the  small  perspective  view. 

On  turning  the  pistons  in  the  directions  shown  in  the  figure, 
water  is  drawn  into  the  space  E  while  water  in  the  space  F  is 


FIG.  10 1. — Behren's  Rotary  Pump. 

being  forced  'out  of  the  pump.  Water  under  pressure  in  F 
cannot  pass  the  surfaces  of  the  pistons,  which  are  always  in 
contact  with  the  fixed  surfaces  of  C  or  D.  It  is  seen  in  the 
figure  that  after  a  little  motion,  the  right-hand  piston  touches 
the  surface  on  the  left-hand  liner  D,  and  then  it  acts  as  an 
abutment  while  the  water  in  the  space  G  is  discharged.  One 
piston  must  come  in  contact  with  the  liner  of  the  opposite  side 
before  the  other  piston  breaks  contact  with  its  opposite  liner. 
In  this  manner  there  is  always  a  barrier  for  a  free  passage  from 
the  discharge  to  the  suction.  This  was  improved  in  the  Port- 
land rotary  pump  of  1882,  when  the  contact  with  the  central  abut- 


114 


PUMPING  MACHINERY 


ment  made  a  line  contact  on  the  center  line,  making  it  unnecess 
ary  to  have  stationary  sleeves,  hence  that  part  was  eliminated. 


t  i  i 

FIG.  102. — MacFarland's  Rotary  Pump. 


The  rotary  pump  of  MacFarland,  1875  (Fig.  102),  represents 
a  development  of  the  Trotter  type  of  rotary  pump.  The  con- 
struction of  this  pump  as  well  as  its  action  can  be  seen  from  the 


FIG.  103. — Phillip's  Rotary  Pump. 

figure.  The  ring  A,  carrying  the  piston  blades  £>,  is  supported 
on  a  stationary  ring  B  projecting  from  the  back  head  of  the 
pump.  The  driving  ring  C  is  carried  from  the  driving  shaft, 
which  is  not  placed  at  the  center  of  the  cylindrical  barrel  of  the 
pump.  The  action  of  the  blades  on  the  water  is  clearly  seen. 


RECENT  HISTORY 


115 


A  pump  somewhat  similar  to  this  was  brought  out 
by  Phillips  about  ten  years  later  (Fig.  103).  In  this 
pump  the  vanes  attached  to  a  sliding  rotating  cylinder 
were  replaced  by  a  cylindrical  roller.  This  arrangement 
would  eliminate  much  of  the  friction,  although,  at  the 
center  and  periphery,  there  is  still  considerable  friction 
produced  by  the  slipping  of  the  rollers  against  the  sides  of 
the  driving  slot. 


FIG.  104. — Wilkin's  Two-Lobed  Rotary  Pump. 


FIG.  105. — Silsby  Rotary  Pump. 

The  pump  (Fig.  104)  of  John  T.  Wilkin,  designed  about 
the  year  1892,  consists  of  two  similar  rotating  lobes  each  having 
two  hypocycloids  and  two  epicycloids.  These  two  curves  will 
work  together  with  uniform  velocity  ratio.  There  is  sliding 
contact  with  the  cylinder  walls  and  at  this  point  there  may  be 
considerable  leakage. 


116 


PUMPING  MACHINERY 


The  use  of  the  curves  employed  for  gear-tooth  profiles  is 
not  uncommon,  as  may  be  seen  in  Fig.  105,  which  represents 
the  form  of  pump  used  on  the  Silsby  fire  engine,  and  in  Fig.  106, 


FIG.  106. — Root  Rotary  Pump. 


FIG.  107. — Allis  Screw  Pump. 

which  represents  the  form  of  propeller  used  by  the  Root  Com- 
pany. These  last  three  pumps  are  modifications  of  the 
early  form  of  Serviere,  illustrated  in  Chapter  I.  In  the 
Root  form  the  number  of  teeth  is  reduced  to  two  for  each  wheel. 
A  development  of  the  early  form  of  screw  pump,  Chapter  I, 
is  shown  in  Fig.  107.  This  pump  is  for  low  lifts,  being  used  in 


RtiCENT  HISTORY 


117 


this  country  when  water  is  to  be  raised  a  few  feet  into  a  stream, 
there  to  mix  with  the  sewage  of  a  town.  The  screw  acts  on 
the  water,  forcing  against  a  static  head.  For  low  lifts  these 


FIG.  108. — Wood  Propeller  Pump. 

wheels,  as  well  as  the  scoop  wheels,  are  quite  effective.     A  devel- 
opment of  the  same  idea  is  shown  in  the  Wood  propeller  pump 


118 


PUMPING    MACHINERY 


(Fig.  108).  In  this  pump  a  vertical  shaft  guided  by  bearings 
00  carries  a  series  of  helical  surfaces  or  screws  NN,  which  fit 
close  to  the  side  of  the  pump  pipe.  Rotating  this  by  means 
of  a  belt,  as  shown  in  the  picture,  the  water  is  forced  upward. 
An  electric  motor  may  be  used  for  driving. 


_FiG.  109. — Helicoidal  Pump. 


FIG.  no. — Bolton  and  Imray  Helical  Pump. 

Wade  &  Cherry's  helicoidal  pump  (Fig.  109)  is  a  combi- 
nation of  the  screw  pump  and  the  centrifugal  pump.  This 
was  brought  out  in  1886.  It  is  so  arranged  that  the  water  is 
first  forced  toward  the  centrifugal  impellers  by  the  screws, 
after  which  it  is  discharged  into  the  central  wheel, 


RECENT  HISTORY 


119 


from  which  place  it  passes  into   the  volute  chamber  around 
the  wheel. 

The  Boulton  &  Imray  helical  pump  (Fig.  no)  was 
described  in  1872.  It  is  an  impeller  pump.  The  casing  con- 
tains a  helical  passage  through  which  the  water  goes  from  one 
level  to  another.  The  height  of  the  passage  is  about  twice  its 
width  and  two-thirds  the  pitch  of  the  helix.  In  this  manner 
it  is  possible  to  place  a  series  of  square  paddles  A  on  a  wheel 
rim  of  such  a  width  that  they  cut  off  one-half  the  height  of  the 
passage,  as  seen  in  the  figure.  The  action  of  the  paddles  is  to 


FIG.  in. — Gould  Fire  Engine. 

impel  the  water  through  the  casing.  The  pump  shown  in  the 
figure  had  a  wheel  3  feet  6  inches  in  diameter  outside  of  the 
blades,  the  blades  themselves  being  6  inches  square.  The 
cylinder  of  the  driving  engine  was  10X8  inches.  Originally 
of  the  form  shown  ft  was  changed  in  a  later  design  by  the  use 
of  the  three-cylinder  Brotherhood  engine. 

The  steam  fire  pump  brought  out  in  the  last  decade  of  the  first 
period  considered  in  this  work  was  improved  during  the  following 
years,  and  by  the  year  1876  a  number  of  them  were  in  use. 


120 


PUMPING   MACHINERY 


The  Gould  steam  fire  pump  (Fig.  in)  exhibited  at  the 
Centennial  Exposition  in  1876  illustrates  to  what  degree  the 
steam  fire  engine  had  developed  from  the  early  form  of  Braith- 
waite  and  Ericsson  described  in  the  last  chapter.  The  boiler 
has  now  been  placed  in  a  vertical  position  and  the  exhaust 
steam  has  been  used  to  produce  a  draft  in  place  of  the  early 
bellows.  The  engine  has  been  made  vertical  also  and  a  fly 
wheel  has  been  added  to  make  the  operation  of  the  pump 
more  regular.  This  type  of  engine  was  likewise  manufactured 
by  Ahrens  and  the  firm  of  Clapp  &  Jones  in  America.  Fig. 


FIG.  112. — Silsby  Fire  Engine. 

112  illustrates  the  application  of  the  rotary  pump  and  engine 
for  fire  service.  The  Silsby  engine  is  one  found  in  many  of 
our  large  cities,  and  although  there  is  considerable  slip  in  the 
pump,  its  light  weight  is  an  important  factor.  These  pumps 
were  introduced  about  the  middle  of  the  last  century.  The 
rotary  steam  engine  is  seen  near  the  boiler  while  the  rotary 
pump  appears  beneath  the  driver's  seat.  The  pump  was  of 
the  type  shown  in  Fig.  105,  the  steam  engine  being  quite  similar. 
The  use  of  horizontal  engines  is  also  found  in  the  types  of 
American  fire  engines,  as  may  be  seen  in  Fig.  113.  The  Button 


RECENT  HISTORY 


121 


engine  is  the  product  of  one  of  the  oldest  builders  of  fire  engines 
and  the  type  is  of  value  in  that  the  parts  are  less  complex  than 
those  in  which  the  fly  wheel  is  used.  The  early  hand  pumps 


FIG.  113, — Button  Fire  Engine. 


FIG.  114. — Hand  Fire  Pump  of  1873. 

were  continued  in  many  places  even  at  this  late  date.  Fig. 
114  illustrates  a  pump  exhibited  at  the  Vienna  Exposition  of 
1873.  This  machine  was  hauled  by  horses,  but  it  was  operated 


122 


PUMPING  MACHINERY 


by  men.  The  handles  A  and  B  are  attached  to  the  pump 
plungers.  The  suction  is  connected  to  the  pipe  or  hose  C  on 
the  rear  of  the  pump,  not  clearly  shown  in  the  figure,  while 
the  discharge  is  delivered  through  D. 

Fig.  115  illustrates  a  modern  type  of  fire. engine  of  recent 
design.  The  pump  is  vertical  with  the  steam  cylinder  near  the 
boiler;  the  pump  end  near  the  point  of  attachment  to  the  fire 
hydrant,  air  chambers  on  the  suction  and  discharge  sides  of 
the  pumps,  valve  boxes  in  accessible  places  and  all  parts  open. 
The  object  sought  in  the  modern  fire  engine  is  a  machine  always 


FIG.  115. — Modern  Fire  Engine,  American-La  France  Co. 

ready  for  a  definite,  positive  service  in  which  all  parts  are 
accessible  for  operation,  maintenance,  and  repair. 

The  air  lift  pump -was  patented  by  James  B.  Frizell  in  1880, 
and  Julius  I.  Pohle  made  a  number  of  improvements  covered  by 
patents  granted  in  1886.  The  pump  (Fig.  116)  consists  of  an 
air  pipe  A  which  dips  into  a  well  B  and  passes  below  the  lower 
end  of  a  larger  pipe  C.  Compressed  air  is  driven  through  the 
pipe  A  and  discharges  against  the  water  pressure  due  to  the 
depth  of  immersion  of  the  discharge  nozzle.  The  air  escapes 


RECENT  HISTORY 


123 


from  the  nozzle  and  acts  on  the  water  above  considered  as  a 
piston,  or  the  air  passes  up  through  the  water,  which  is  thus 
made  lighter  and  is  forced  upward  by  the  static  pressure  of 
the  water  in  the  well.  The  figure  shows  beads  of  water  acting 
as  pistons  with  air  between,  expanding  as  the  head  is  reduced 
by  the  upward  motion  of  the  water.  This  is  probably  the 


FIG.  1 1 6. — Air-Lift  Pump. 

manner  of  action  of  the  pump  after  is  it  in  operation,  although 
in  starting  it  is  likely  that  the  reduction  of  density  by  aeration 
causes  the  first  flow.  Although  this  form  of  pumping  apparatus 
was  not  extensively  used  or  seriously  thought  of  before  the 
work  of  Pohle  in  1886,  the  method  had  been  proposed  earlier. 
Collom's  Lectures  on  Mining  for  1876,  delivered  in  Paris, 
described  this,  and  Gerlach  in  a  paper  points  out  that  Loescher 


124 


PUMPING  MACHINERY 


of  Freiberg  described  a  similar  apparatus  in  a  pamphlet  printed 
in  1797. 

Another  form  of  pumping  apparatus  using  air  is  that  in- 
vented by  Professor  Elmo  G.  Harris  about  1900.  In  this  pump 
(Fig.  117)  air  is  compressed  by  the  compressor  A  into  a  pipe 


FIG.  117. — Harris  Air-Lift  Pump. 

leading  to  a  tank  B.  This  compressed  air  acts  on  top  of  the 
water  contained  in  B  and  drives  it  out  of  the  discharge  pipe  D. 
The  suction  side  of  the  compressor  is  connected  with  a  pipe 
leading  to  the  tank  C  and  the  reduction  of  pressure  within  the 
tank  draws  water  into  C.  At  the  proper  time  the  valve  E 
is  shifted,  changing  the  suction  and  pressure  sides  so  that  water 


RECENT  HISTORY 


125 


is  sucked  into  B  and  driven  from  C.  There  is  an  equalization 
of  pressure  through  the  valves  of  the  compressor  as  soon  as 
the  shifting  valve  E  is  turned  and  before  the  compressor  starts 
to  drive  out  the  water  from  C.  This  utilizes  part  of  the  energy 
used  in  compressing  the  air  into  cylinder  B. 

These  two  air  lift  pumps  of  Pohle  and  Harris  have  the 
advantage  that  a  number  of  them  may  be  driven  from  a  central 
plant  and  there  is  little  loss  from  such  an  arrangement.  The 
air  lift  pump  raises  a  larger  quantity  in  a  given  time  than 
would  be  possible  with  a  deep  well  pump.  The  Harris  pump 
may  be  installed  in  places  where  other  pumps  are  impossible, 


FIG.  118. — Giffard  Injector. 

and  forms  a  positive  type  of  pump.  These  pumps  have  specific 
advantages  which  will  be  pointed  out  when  their  design  is 
considered. 

In  1858  Henri  Jacques  Giffard  patented  his  invention  ,of 
the  steam  injector — one  of  the  simplest  devices  for  pumping 
water  into  boilers.  In  the  earliest  form  (Fig.  118)  steam 
entered  through  the  cock  A  and  passed  through  a  series  of 
small  holes  into  the  interior  of  a  tube  having  a  nozzle  B  at 
its  extremity.  The  core  or  rod  C  controlled  by  D  closes  the 
nozzle  or  regulates  the  amount  of  opening.  The  steam  acquires 
a  high  velocity  in  the  nozzle,  causing  a  vacuum  in  the  space 
E,  which  raises  water  to  that  point,  where  it  mixes  with  the 
steam.  As  the  steam  condenses  and  imparts  a  high  velocity  to 
the  water,  the  cross-section  of  the  mixture  grows  smaller  and 


126  PUMPING  MACHINERY 

consequently  the  combining  tube  F  is  made  convergent.  To 
change  this  high  velocity  into  pressure,  Giffard  followed  the 
converging  combining  tube  by  a  diverging  delivery  tube  G.  In 
this  manner  the  velocity  was  reduced  and  the  pressure  was 
so  increased  that  the  water  could  enter  against  the  pressure  of 
the  boiler. 

The  idea  of  using  the  power  of  a  moving  jet  was  not  new. 
According  to  Kneass  a  crude  injecting  apparatus  was  used  as 
early  as  1570  by  Vitrio  and  Philibert  de  Lorme.  In  1818 
Mannoury  d'Ectot  patented  a  device  which  was  quite  similar 
to  the  injector,  and  Bourdon  in  1857  patented  a  combination 
of  convergent  and  divergent  tubes  for  transforming  the  energy 
of  a  moving  jet. 

The  latter  invention  resembled  Giffard' s,  but  the  writings 
of  the  inventor  of  the  injector,  made  seven  years  before  Bour- 
don's invention,  in  which  he  explained  the  theory  of  the  injector, 
proved  that  he  should  receive  the  credit  for  this  invention. 

The  manufacture  of  these  machines  was  at  once  undertaken 
in  various  countries  by  licensees.  In  America  William  Sellers 
&  Co.  began  their  manufacture,  while  Sharp,  Stewart  &  Co., 
made  them  in  England. 

The  Giffard  design  was  an  excellent  one.  Many  ideas 
were  suggested  in  his  patent  specifications  and  pamphlets, 
so  that  although  changes  were  made  in  the  injector  as  manu- 
facturing processes  were  developed,  and  as  more  knowledge 
was  had  of  the  properties  of  jets  and  steam,  the  original  ideas 
are  still  to  be  seen  in  the  present-day  injector. 

The  operation  and  design  of  injectors  will  be  considered 
later  with  the  modern  form  of  injector,  yet  it  is  well  to  men- 
tion the  names  of  some  of  those  who  added  to  the  development 
of  the  injector.  They  are:  Millholland,  Rue,  Sellers,  Han- 
cock, Williams,  Loftus,  Bancroft,  and  Kneass  in  America; 
Robinson  &  Gresham,  Turck,  Hamer,  Metcalf,  and  Davies  in 
England;  Haswell,  Pradel  &  Krauss  and  Korting  in  Germany; 
Cuau,  Bouvret,  Polonceau,  and  Delpeche  in  France. 

A  pump  in  which  the  driving  steam  touches  the  water  is 
known  as  the  pulsometer  (Fig.  119).  This  is  really  a  develop- 


RECENT  HISTORY 


127 


ment  of  the  old  Savery  engine.     Steam  is  admitted  at  A  and 

if  the  ball  is  resting  over  the  left-hand  opening  steam  will 

enter    C    and    drive    out    the 

water.     A  sudden  change  in  the 

steam  pressure  when  the  water 

surface  reaches  the  level  of  the 

discharge  valv$  causes  the  ball 

to    close    the    right-hand    side, 

and  then  the  condensation  of 

the  steam  on  that  side  draws 

water   through   the   suction  D 

and  suction  valve  E,  as  shown 

in  the  figure. 

When  the  water  is  driven 
from  B  the  ball  is  forced  over 
to  the  left  and  the  operation  is 
repeated. 

This  type  of  pump  is  easily 
applied,  and  in  contracting  work 

it      is      USed      for      this      reason,  FIG.  ng.-Pulsometer. 

although  it  is  most  expensive 

when  its  thermal  efficiency  is  considered.  For  this  kind  of 
rough  and  rapid  work  it  may  not  cost  more  in  money  to  operate 
this  than  other  forms  of  pumps. 


FIG.  1 20. — Quimby  Screw  Pump. 


128 


PUMPING   MACHINERY 


The  Quimby  pump  (Fig.  120)  is  a  development  of  the  pump 
shown  in  Chapter  I,  as  the  invention  of  Revillion.  Its  oper- 
ation is  evident  from  the  figure,  which  shows  right-  and  left- 
handed  screws  meshing^  together  so  as  to  form  abutments  for 
each  other,  catching  the  liquid  in  the  threads,  thus  forcing  it 
onward. 

The  latest  form  of  pumping  apparatus  designed  is  that 
of  Herbert  A.  Humphrey,  described  by  him  in  Engineering, 
November  26,  1909.  It  is  a  combined  gas  engine  and  pump. 
The  pump  (Fig.  121)  consists  of  an  explosion  head  A,  a  suction 


FIG.  121. — Explosion  Pump  of  Humphrey. 

portion  B,  a  discharge  pipe  C.  Assume  that  a  compressed 
charge  of  gas  and  air  is  exploded  in  the  head  A.  The  force 
of  this  combustion  drives  the  water  from  C  into  the  reservoir, 
and  the  energy  given  to  the  water  by  the  excess  pressure  in  A 
is  dissipated  only  after  the  pressure  in  A  has  been  reduced  to  a 
vacuum.  When  the  pressure  at  B  within  the  pipe  system  is 
reduced  sufficiently,  water  is  drawn  in  through  the  suction 
valves.  After  the  water  is  brought  to  rest  in  C,  the  gases  in  A 
are  compressed  by  the  static  pressure  from  the  reservoir,  which, 
forces  the  water  backward,  developing  a  certain  amount  of 
velocity.  During  this  operation,  however,  the  valve  D  is 
opened  automatically,  and  the  burned  gases  are  exhausted. 


RECENT  HISTORY 


129 


When  the  water  reaches  a  point  near  the  valve  D  the  valve 
is  closed,  and  before  the  velocity  which  has  been  set  up  in  C 
can  be  dissipated  the  water  and  gas  in  A  are  compressed  beyond 
the  pressure  corresponding  to  the  static  head.  This  means 
that  there  is  another  surge  toward  the  reservoir,  followed  by  a 
vacuum  in  A,  which  is  destroyed  by  opening  the  air  and  gas 
valve  E,  permitting  a  fresh  charge  to  enter.  This  suction 
stroke  is  followed  by  another  backward  surge,  during  which 
the  explosive  mixture  in  A  is  compressed.  At  the  proper 
time  this  is  exploded  by  an  electric  spark,  and  the  operation 
is  repeated.  The  valves  E  and  D  and  the  ignition  apparatus 
are  operated  by  the  surging  water  in  the  system.  The  device 
is  so  arranged  that  E  opens  on  an  expansion  stroke  after  D 
closes  at  the  end  of  a  compression  stroke. 

Professor  W.  C.  Unwin  made  a  test  on  this  pump  in  connec- 
tion with  a  gas  producer,  developing  16  pump  horse  power  on 
1.063  pounds  coal  per  delivered  horse  power  hour  or  12,243 
B.T.U.  per  P.H.P.  hour.  The  data  for  this  test  are  given  in 
the  table  below: 


Lift. 

P.H.P. 

Cu.  ft.  at 
14.7-32°  P. 

Calorific 
Value. 

B.T.U. 
per  P.H.P.  hr. 

Lbs.  Anth. 
per  P.H.P.  hr. 

32.87 

16.15 

83.12 

!47-3 

12,243 

I  .063 

2S-95 
20.73 

12.32 
10.99 

90.93 
93.61 

*43-5 
I4S-3 

!3.°37 
*3,596 

1.132 
1.180 

Mr.  Humphrey  does  not  claim  to  be  the  originator  of  a  gas- 
driven  pump,  as  he  says  this  matter  dates  back  to  1868,  but  he 
deserves  credit  for  having  built  so  simple  a  device  which  yet 
gives  efficiencies  higher  than  those  of  the  most  improved  steam 
pumps. 


CHAPTER  III 
MODERN  RECIPROCATING   PUMPS 

IN  considering  the  actual  pumps  in  use  it  is  advisable  to 
classify  them  in  several  ways:  (a)  In  regard  to  the  form  of 
water  displacer;  (b)  in  regard  to  the  number  of  displacements; 
(c)  in  regard  to  the  method  of  operation;  (d)  in  regard  to  the 
manner  of  packing;  (e)  in  regard  to  the  direction  of  the  axis 
of  the  pump  cylinder;  (/)  in  regard  to  the  arrangement  of  cyl- 
inders; (g)  in  regard  to  the  use  of  the  pump. 

The  general  forms  of  that  part  of  a  pump  used  to  displace 
water  are  the  plunger,  the  piston,  and  the  bucket,  although 
such  means  as  air,  gas,  and  steam  are  used,  as  was  shown  in 
Chapters  I  and  II.  The  plunger  is  usually  cylindrical  in  section 
and  forces  the  water  of  a  pump  by  entering  the  space  occupied 
by  the  water.  Fig.  122  shows  the  construction  of  a  simple 
plunger  pump.  This  plunger  enters  the  cavity  filled  with 
water  and  displaces  an  amount  equal  to  the  increase  of  volume 
of  the  plunger  protruding  into  the  cylinder. 

The  piston  (Fig.  123)  consists  of  a  movable  diaphragm 
tightly  fitting  against  the  sides  of  a  cylinder,  and  forcing  the 
water  before  it.  This  piston  forces  an  amount  of  water  equal 
to  the  area  of  the  piston  multiplied  by  the  stroke. 

The  bucket  is  a  piston  containing  a  number  of  valve-closed 
passages  through  which  the  fluid  may  pass  at  the  proper  time. 
Fig.  124  shows  the  construction  of  a  bucket. 

The  use  of  any  one  of  these  devices  on  a  pump  gives  its 
name  to  the  pump.  Thus  the  names  plunger  pump,  piston 
pump,  or  bucket  pump  designate  a  pump  in  which  one  of  these 
displacers  is  used. 

Single  plungers  and  buckets,  as  ordinarily  constructed,  can 
lift  water  on  only  one  stroke,  while  the  piston  displaces  water 

130 


MODERN  RECIPROCATING  PUMPS 


131 


on  both  strokes.     This  gives  rise  to  the  classes,  single-  and 
double-acting  pumps.     Of  course  two  plungers  may  be  united 


O 


LJ 


PQ 


IT 


o 


by  an  outside  or  an  inside  connection  and  form  a  double-acting 
plunger  pump  as  shown  in  Fig.  125,  while  a  bucket  or  plunger 
pump  may  be  so  arranged  that  although  it  lifts  only  on  one 


132 


PUMPING  MACHINERY- 


stroke  of  every  two  strokes,  it  discharges  on  each  stroke.  Such 
pumps  are  known  as  differential  pumps.  Fig.  126  shows  the 
arrangement  for  a  plunger,  and  the  same  could  be  used  with  a 
bucket  (Fig.  127).  The  piston  rod  is  these  cases  is  so  enlarged 
that  it  is  practically  another  plunger  of  about  one-half  the  area 


f^^s^fv  ^"^ 

D 

D 

^"'"^\^S?1^ 

[] 

n 
u 



FIG.  125. — Double-Acting  Plunger  Pump. 

of  the  main  plunger  or  bucket.  In  this  manner  by  connecting 
the  lifting  side  of  the  pump  to  the  discharge  main  through  the 
other  end  of  the  pump,  one-half  of  the  water  will  be  retained 
on  this  side  during  'the  discharge  stroke  of  the  main  plunger, 
while  one-half  of  the  water  is  discharged.  The  portion  of  the 


FIG.  126. — Differential  Plunger  Pump. 


FIG.  127. — Differential 
Bucket  Pump. 


water  retained  is  delivered  on  the  suction  stroke  of  the  main 
plunger  or  bucket.  In  this  manner  the  discharge  is  made  more 
regular,  although  the  suction  occurs  on  every  other  stroke. 

The  method  of  operation  gives  another  classification. 
Steam  pumps  are  those  driven  by  steam  pistons,  while  com- 
pressed-air pumps  are  those  in  which  air  is  used  in  place  of 


MODERN  RECIPROCATING  PUMPS 


133 


steam.  Where  the  steam  and  water  pistons  are  directly  con- 
nected these  are  known  as  direct  acting,  although  if  a  fly  wheel 
is  connected  to  the  system  they  are  known  as  fly-wheel  pumps. 


FIG.  128. — Plunger  and  Ring  Packing. 

Power  pumps  are  those  driven  by  belts  or  gearing,  and  electric 
pumps  are  those  directly  driven  by  electric  motors.  The  use 
of  air  or  gas  in  direct  contact  with  the  water  has  given  the 
class  of  air-lift  and  gas-driven  pumps  described  in  Chapter  II, 


FIG.  1 29. — Central  Outside  Packing. 

while  in  the  injector  and  pulsometer  the  water  is  moved  by 
the  direct  action  of  steam. 

The  classification  due  to  the  method  of  packing  the  water 
piston  or  plunger  divides  pumps  into  outside- packed  pumps 
(Fig.  125),  and  inside-packed  pumps  (Fig.  128),  as  well  as  central- 


134  PUMPING  MACHINERY 

and  end-packed  machines.  A  special  type  of  inside  packing 
is  known  as  the  plunger  and  ring  type.  Fig.  128  shows  the 
arrangement  of  this  type  of  packing.  The  plunger  in  this  case 
passes  through  a  long  sleeve  or  ring  in  which  the  resistance 
against  the  flow  of  water  is  so  great  that  there  is  little  or  no 
leakage.  Fig.  125  shows  an  end  packing  while  Fig.  129  shows 
central  packing. 

Vertical  and  horizontal  pumps  are  distinguished  by  the 
direction  of  the  axis  of  the  pump. 

A  pump  with  a  single  water  cylinder  is  called  a  simplex 
pump,  while  two-cylinder  pumps  are  called  duplex,  and  three- 
cylinder,  triplex.  The  first  two  names,  however,  are  usually 
associated  with  direct-acting  pumps  of  small  or  medium  size, 
while  the  last  term  is  usually  applied  to  a  type  of  single-acting 
power  pump  in  which  there  are  three  cylinders. 

The  last  classification  of  pumps  is  that  due  to  the  use  of  the 
pump.  It  is  one  of  the  largest  classifications  and  will  demand 
a  more  detailed  consideration.  In  examining  these  different 
classes  the  peculiar  features  of  e.ach  will  be  pointed  out.  For 
those  types  which  will  not  be  considered  later  in  this  work, 
a  full  description  will  be  given  here. 

Boiler-Feed  Pumps.  This  type  of  pump  is,  in  all  probability, 
the  most  common.  The  conditions  under  which  such  pumps 
are  used  determine  many  of  their  details.  In  general  the  pump 
has  a  water  piston  or  plunger  which  is  about  one-third  the  size 
of  the  steam  cylinder.  The  object  of  such  a  difference  is  to 
have  ample  force  with  any  amount  of  steam  pressure  to  drive 
water  into  the  boiler  against  the  steam  pressure  which  operates 
the  pump.  The  pump  must  be  as  simple  as  possible,  since  it  is 
to  be  handled  by  unskilled  men.  It  must  have  few  parts 
which  are  liable  to  breakage  or  disarrangement.  As  the  varia- 
tion in  pressure  in  the  discharge  line  is  not  important  and  as  the 
pumps  are  rarely  run  at  full  speed,  it  is  quite  customary  to 
install  these  pumps  without  air  chambers  on  the  discharga 
or  suction. 

These  pumps  are  either  simplex  or  duplex,  as  may  be  seen 
from  Fig.  130  and  Fig.  131.  Fig.  130  shows  two  small  Worth- 


MODERN  RECIPROCATING  PUMPS 


135 


ington  pumps,  the  upper  one  being  a  piston  pump  and  the 
lower  an  end  outside-packed  plunger  pump.  These  are  both  of 
the  same  size,  6  inches  diameter  of  steam  cylinder,  4  inches 
diameter  of  water  cylinder,  and  a  common  stroke  of  6  inches. 
This  is  usually  written  as  6x4X6  inches.  Fig.  131  shows  a 


FIG.  130. — Duplex  Boiler  Feed  Pumps. 
(Sizes  6X4X6.)     , 

10x6x12  inch  Knowles  simplex  boiler-feed  pump.     The  water 
end  of  this  pump  is  of  the  piston  type. 

For  heavier  water  pressures  the  boiler-feed  pump  is  made 
with  steam  cylinders,  large  when  compared  with  the  water 
end,  and  the  valves  are  placed  above  the  cylinder  casting  in 
separate  valve  pots  as  shown  in  Fig.  132.  This  pump,  12  inches 


136 


PUMPING  MACHINERY 


and  17X10X15  inches,  is  a  compound  pump  in  which  two 
steam  cylinders  are  used  on  each  side  of  the  duplex  pump  for 


FIG.  131. — Simplex  Boiler  Feed  Pump. 
(Size  10X6X12.) 

the  purpose  of  getting  greater  economy.      Fig.  133  shows  a 
different  design  for  the   same   type   of  pump.     The   outside 


FIG.  132. — Compound  Outside  Center  Packed  Boiler  Feed  Pump. 
(Size  12  and  17X10X15.) 

packing  in  this  case  is  of  the  end  type  instead  of  the  center 
type  shown  in  Fig.  132. 

For  use  in  marine  installations  where  floor  space  is  valuable 


FIG.  133.— Outside  Packed  Plunger  Boiler  Feed  Pump. 


FIG.  134. — Marine  Boiler  Feed  Pump. 
(Size  10X7X12.) 


137 


138 


PUMPING  MACHINERY 


vertical  pumps  are  designed.  These  are  often  known  as 
Admiralty  Pumps.  Fig.  134  shows  a  Davidson  duplex 
piston  pump  bolted,  to  a  bulkhead  of  a  vessel.  It  is  to  be 
noted  that  this  pump  has  its  valves  so  placed  that  they  may 
be  examined  readily.  The  suction  and  discharge  pipes  may 
be  attached  to  either  side,  and  caps  on  the  cylinder  heads 
allow  an  examination  of  the  water  piston. 

To  give  some  idea  of  sizes  of  boiler-feed  pumps  the  following 
tables  have  been  taken  from  catalogues  of  pump  makers: 

THE    WORTHINGTON    BOILER-FEED    PUMP 
Pressure  Pattern — For  250  Pounds  Pressure. 

These  pumps  have  four  single-acting,  outside-packed  water 
plungers,  working  through  adjustable  stuffing  boxes  in  the 
ends  of  the  water  cylinders.  The  valves  are  of  brass,  guided 
from  below  by  wings  and  controlled  by  composition  springs, 
and  are  located  in  separate  valve  chambers  or  pots  designed 
to  withstand  the  heavy  pressures  to  which  this  pump  may  be 
subjected. 


Diam- 
eter of 

Diam- 
eter of 

Length 

e> 

H.P.  Boiler, 
Based  on  45 
Ibs.  Water  pet 

Sizes  of  Pipes  for  Short  Lengths 
to  be  Increased  as  Length 
Increases. 

Approximate 
Space  Occupied, 
Feet  and  Inches. 

Steam 

Water 

of 

hour,   which 

Cylin- 
ders. 

Plun- 
gers. 

Stroke. 

Pump  will 
Supply  at 
Slow  Speed. 

Steam 
Pipe. 

Ex- 
haust 
Pipe. 

Suc- 
tion 
Pipe. 

De- 
livery 
Pipe. 

Length. 

Width. 

4* 

2 

4 

70 

* 

! 

Il 

I 

3      91 

i      3 

si 

3 

5 

190 

i 

i* 

al 

2 

4    ii 

!        6§ 

6 

Si 

6 

290 

I 

it 

2* 

2 

5     7 

i      6* 

71 

4* 

6 

470 

Ii 

2 

,4 

3 

5  10 

2        I 

71 

4i 

10 

670 

ii 

2 

4 

3 

8     3 

2        I 

9 

5 

10 

800 

2 

»i 

4 

3 

8     3 

2        I 

10 

6 

10 

I2OO 

2 

*i 

6 

5 

8   10 

4        2 

12 

7i 

IO 

I4OO 

«1 

3 

6 

5 

9    ii 

4     3 

'14 

*| 

10 

I800 

*i 

3 

8- 

7 

IO        I 

3      3 

12 

71 

*5 

20OO 

*i 

3 

6- 

5 

ii      9 

4     3 

1  14 

81 

IS 

2700 

*i 

3 

8 

7 

II        2 

4     o 

'17 

10 

IS 

3700 

2* 

3i 

8 

7 

12        2 

4     3 

1   20 

12 

15 

5200 

4 

5 

IO 

8 

M     5 

4     5 

These  sizes  have  four  center-packed  plungers. 


MODERN  RECIPROCATING  PUMPS 


139 


PRESCOTT  DUPLEX   OUTSIDE-PACKED    PLUNGER  POT-FORM 
BOILER-FEED    PUMPS 

For  300  Pounds  Working  Pressure. 

The  water  ends  are  of  the  "  pot  form,"  and  have  four 
single-acting  outside-packed  water  plungers.  The  water  valves 
are  designed  for  either  hot  or  cold  water  and  arranged  so  as  to 
be  readily  accessible.  All  water  passages  are  large  and  direct. 

The  plunger  stuffing  boxes  are  very  deep  and  fitted  at  the 
bottom  with  a  removable  brass  ring  which  can  be  replaced 
when  worn. 

These  pumps  are  especially  adapted  for  boiler-feeding  service 
in  electric  lighting  and  railway  stations  or  in  other  plants 
where  high  pressures  are  carried. 


Size. 

(U"-1   O 

Diameter  of  Pipe  Openings. 

Space  Occupied, 
Feet  and  Inches. 

^ 

S^e* 

, 

1 

fe 

jj 

«  |"0ffi 

*j 

c 

I 

SJ 

1 

a 

rt 

1 

Length. 

Width. 

z 

£* 

£ 

GQ 

•  w 

s 

8 

4 

12 

500 

Ii 

2 

3i 

3 

9     9i 

4      6J 

10 

5 

12 

800 

2 

a* 

4 

2} 

IO        2^ 

4      6J 

12 

6 

12 

1,200 

2 

3 

6 

5 

10        34 

5     9 

12 

7 

12 

2,000 

2 

3 

6 

,  5 

ii      4 

5     5 

14 

8 

18 

4,3°° 

a| 

3} 

7 

6 

J4     3 

4     3 

16 

10 

18 

6,850 

2| 

si 

8 

6 

14    si 

4      ii 

18 

10 

18 

6,850 

3 

4 

8 

6 

M     7* 

4      ii 

24 

14 

24 

25,000 

5 

6 

12 

10 

19     oj 

5    "4 

THE   DAVIDSON    VERTICAL   DUPLEX    PUMP 
For  a  Pressure  of  250  Pounds. 

In  the  Davidson  vertical  duplex  pump  there  are  but  few 
joints,  all  of  which  are  visible  and  easily  renewed,  the  water 
valves  are  easily  accessible  for  examination  or  renewal,  the  water 
pistons  can  be  packed  from  the  upper  end  of  cylinder  and  are 
fitted  with  fibrous  or  metallic  packing.  Water  ends  of  cast 
iron  are  composition  lined  and  fitted,  or  entirely  of  composi- 
tion when  specially  ordered. 


140 


PUMPING  MACHINERY 


H.  P.  Boiler, 

Steam 
.Cyl- 
inder. 

Water 
Cyl- 
inder. 

Stroke, 
Inches. 

Gallons 
per  Single 
Stroke  of 
each 

based  on  30  Ibs. 
of  Water  oer 
H.  P.  per-  Hour, 
which  the 

Steam 
Pipe. 

Exhaust 
Pipe. 

Suction 
Pipe. 

Dis- 
charge 
Pipe. 

Piston. 

Pump  will  sup- 

ply with  Ease. 

4 

*1 

4 

.084 

165 

| 

1 

2 

Ii 

4* 

2| 

6 

•154 

300 

\ 

22 

2 

5i 

3i 

6 

•25 

500 

i 

Ji 

3 

ai 

6 

4 

8 

•435 

870 

i 

ii 

3 

M 

7 

4 

8 

•435 

870 

i  \ 

i\ 

3 

24 

7 

4i 

8 

•55 

IIOO 

^ 

Ii 

4 

3 

8 

5 

10 

•    -85 

1700 

ii 

2 

4 

si 

8 

5 

12 

i  .02 

200O 

ii[ 

2 

4 

9 

5i 

IO 

1.03 

2OOO 

if! 

2 

4^ 

4 

10 

6 

10 

1.225 

2450 

2 

21 

*5 

41 

10 

6 

12 

i  .469 

2900 

2 

2* 

5 

4i 

12 

7 

12 

2  .OO 

4OOO 

2 

a| 

6 

5 

14 

8 

12 

2.6l 

500O 

2| 

3 

7 

6 

14 

8i 

12 

2.94 

6OOO 

24      ' 

3 

7 

6 

Suction  and  discharge  openings  on  both  sides. 

Capacities  for  boiler  feeding  are  based  on  a  speed  of  60  single 


FIG.  135. — Worthington  Packed-Plunger  Pump. 
(Size  7*X5X~6.) 

strokes  a  side  per  minute;    for  other  services,  pumps  should 
be  run  at  a  piston  speed  of  30  to  80  feet  a  side  per  minute,  and 


MODERN  RECIPROCATING  PUMPS  141 

in  cases  of  emergency  can  be  speeded  up  greatly  in  excess 
of  this. 

General  Service  Pumps  are  those  which  are  intended  for 
the  pumping  of  water  for  various  purposes;  drainage,  elevator 
work,  small  water  supply  or  any  other  work  of  a  general  nature. 
These  pumps  are  usually  designed  for  specific  pressures.  For 
water  pressures  of  200,  250,  and  300  pounds  per  square  inch  the 
outside  packed  plunger  water  ends  (Fig.  135 )  are  used,  while 


FIG.  136. — Tank  Pump. 
(Size  12X15X15.) 

with  pressures  of  about  150  pounds  piston  pumps  are  used. 
When  the  water  pressure  is  from  35  to  50  pound  the  pump  is 
known  as  a  tank  piimp,  and  although  the  general  form  is  the 
same  as  that  of  the  other  pumps,  the  parts  are  made  lighter 
and  the  steam  cylinder  is  made  smaller  than  the  water  end. 
In  all  cases  of  these  pumps  the  variation  from  one  class  to 
another  depends  on  the  pressure  to  be  carried.  Fig.  136  shows 
the  type  of  tank  pump,  while  the  following  table  gives  the 
sizes  in  use  by  one  manufacturer. 


142 


PUMPING  MACHINERY 


THE    WORTHINGTON    PISTON    PUMP 
For  Tank  or  Light  Service. 


Diameter  of 
Steam  Cylinders. 

Diameter  of 
Water  Pistons. 

Length  of  Stroke. 

Gallons  per 
Revolution. 

Maximum  Revo- 
lutions per 
Minute. 

Maximum  Gallons 
per  Minute. 

Sizes  of  Pipes  for  Short 
Lengths  to  be  increased 
as  Length  increases. 

Approximate 
Space  Occupied, 
Feet  and  Inches. 

§| 

JE 

09 

Exhaust 
Pipe. 

IJ 

"u£ 
3 
O! 

U 

r 

Length. 

Width. 

3 

2| 

3 

•3 

80 

24 

i 

i 

It 

! 

2    io£ 

o      9} 

4i 

3i 

4 

•75 

75 

56 

i 

1 

2i 

xi 

2     II 

i      i 

Si 

4l 

5 

i-.S1 

70 

106 

i 

1} 

-  3 

2 

3      3 

i      4 

6 

si 

6 

2.65 

65 

172 

i 

1} 

4 

3 

3    I0 

i      5 

7i 

5l 

6 

2.65 

65 

172 

i| 

2 

4 

3 

3    JI 

I     10 

6 

7i 

6 

4-54 

65 

295 

i 

1} 

6 

5 

3      9 

i      9 

7i 

7| 

6 

4-54 

65 

295 

i* 

2 

6 

5 

3      9 

i      9 

6 

*f 

6 

5-84 

65 

380 

i 

li 

6 

5 

3      9 

I     IO 

ri 

8} 

6 

5-84 

65 

380 

a 

2 

6 

5 

3      9 

I     IO 

7i 

6 

10 

4-75 

54 

256 

ii 

2 

5 

4 

5      8 

2      4 

7i 

7 

IO 

6.52- 

54 

'  352 

»i 

2 

6 

5 

5      8* 

2      4 

IO 

7 

IO 

6.52 

54 

352 

2 

2i 

6 

5 

5     9 

2     5 

7i 

8i 

IO 

9.68 

54 

522 

1| 

2 

6 

5 

6     o 

2      4 

9 

8J 

IO 

9.68 

54 

522 

2 

2* 

6 

5 

6     o 

2      4 

71 

roi 

10 

14.08 

54 

760 

I* 

2 

IO 

8 

6     4 

2        8 

9 

10} 

IO 

14.08 

54 

760 

2 

•1 

IO 

8 

6      5 

2        8 

12 

ioi 

IO 

14.08 

54 

760 

2i 

3 

10 

8 

6     9 

2        8 

12 

12 

IO 

J9-37 

54 

1045 

•1 

3 

IO 

8 

6      5 

3      o 

12 

14 

10 

26.44 

54 

1427 

2* 

3 

12 

10 

7     o 

3      4 

12 

IS 

IO 

30.38 

54 

1640 

2* 

3 

12 

10 

7     o 

3      4 

14 

15 

10 

30.38 

54 

1640 

2\ 

3 

12 

IO 

7     o 

3      4 

12 

14 

15 

39-70 

40 

1588 

^ 

3 

12 

10 

7    10 

3    10 

12 

^5 

15 

45-6i 

40 

1824 

M 

3 

12 

IO 

7    10 

3    10 

14 

IS 

15 

45-6i 

40 

1824 

ai 

3 

12 

IO 

7    10 

3    10 

17 

IS 

15 

45-61 

40 

1824 

2i 

3i 

12 

IO 

8     o 

3    10 

12 

J7 

15 

58.66 

40 

2346 

ai 

3 

14 

12 

8     5 

4      i 

14 

i7 

15 

58.66 

40 

2346 

ti 

3 

14 

12 

8     5 

4      i 

14 

19 

15 

73-27 

40 

2931 

2* 

3 

16 

14 

8     o 

4      i 

17 

J9 

15 

73-27 

40 

2931 

2i 

3i 

16 

14 

8      ii 

4      i 

14 

22 

15 

92.26 

40 

3940 

2i 

3 

16 

14 

8     6 

4      7 

An  additional .  charge  is  made  for  Tobin-bronze  piston  rods,   brass 
water  pistons,  bed  plates,  or  for  any  extras. 

The  water  end  is  made  of  light  construction  for  use  on 


MODERN  RECIPROCATING  PUMPS  143 

services  where  the  total  water  pressure  to  be  pumped  against 
is  not  over  from  35  to  50  pounds  per  square  inch.  The  ratios 
of  steam  cylinders  and  water  pistons  are  suitable  for  raising 
liquids  to  moderate  heights  with  ordinary  steam  pressures. 
This  design  is  intended  for  use  at  railway  water  stations, 
breweries,  distilleries,  gas  and  oil  works,  tanneries,  bleacheries, 
refineries,  etc.  The  valves  in  the  liquid  end  are  furnished  of 
material  suitable  for  the  liquid  to  be  pumped.  The  liquid 
ends  up  to  the  size  14  X 10  are  designed  for  pressures  up  to  50 
pounds  per  square  inch  and  the  larger  ends  for  a  pressure  of  35 
pounds  per  square  inch. 

A  tank  pump  used  for  pumping  water  from  the  bottom  of 
a  vessel  is  known  as  a  bilge  pump  or  ballast  pump.  These 
pumps  are  usually  placed  in  a  vertical  position  owing  to  the 
lack  of  room.  On  account  of  the  small  head  against  which 
they  work  the  steam  cylinder  is  much  smaller  than  the  water 
cylinder.  Fig.  137  shows  the  form  of  ballast  pump  built  by 
Worthington  and  the  covers  over  the  valve  chambers  show 
how  simple  it  is  to  care  for  the  valves.  This  is  an  important 
matter  in  pumps  for  marine  service  where  working  space  is 
limited. 

Fire  Pumps  are  those  installed  for  fire  protection.  When 
these  are  built  according  to  certain  specifications  adopted  by 
the  Underwriters  Associations  they  are  known  as  under- 
writers' pumps.  Fig.  138  shows  such  a  pump  with  ample 
dimensions  on  the  steam  end.  As  will  be  seen  later  these 
pumps  have  gauges  on  the  steam  and  water  ends,  air  chambers 
on  the  suction  and  discharge  pipes,  direct  hose  connections  on 
the  pumps,  a  relief  valve  on  the  discharge  main,  and  a  name 
plate  with  certain  data  in  regard  to  the  pump  on  the  air 
chamber.  The  water  parts  are  of  bronze  and  there  are  other 
peculiar  features  which  will  be  considered. 

THE    KNOWLES    UNDERWRITER    FIRE    PUMPS 
1904  Pattern 

In  these  underwriter  fire  pumps  the  water  passages,  valve 
areas,  and  suction  and  discharge  nozzles  are  all  much  larger 


144 


PUMPING  MACHINERY 


than  in  any  ordinary  pump,  so  a  greater  amount  of  water  can 
be  discharged  without  water  hammer.  Also,  the  steam  and 
exhaust  ports  and  nozzles  are  designed  so  as  to  give  unre- 
stricted passage  to  the  steam.  The  pump  is  "  rust  proof," 


FIG.  137. — Worthington  Ballast  Pump. 
(Size  7iXioiXio.) 

and  will  start  instantly,  after  standing  unused  for  a  long  time. 
The  piston  rods  and  valve  rods  are  made  from  Tobin  bronze. 
All  stuffing  boxes  and  glands  are  brass  lined.  Plungers  and 
plunger  sleeves  are  of  composition,  but  the  metals  are  differ- 
ently mixed  to  prevent  cutting.  This  mixture  of  the  metals 
has  been  a  subject  of  much  experiment  and  study;  it  has  now 


MODERN  RECIPROCATING  PUMPS 


145 


FIG.  138. — Knowles  Underwriters'  Pump. 
(Size  18X10X12;  1000  gallons  per  minute.) 


FIG.  139. — Compound  Outside  End-Packed  Pressure  Pump. 
(Size  10X16X4X21.) 


146 


PUMPING  MACHINERY 


brought  great  success,  as  the  parts  work  perfectly  upon  each 
other,  without  friction  or  impairment. 

Each  pump  has  the  following  fittings: 

Capacity  plate  on  discharge  air  chamber;  stroke  gauge, 
graduated  on  each  end;  vacuum,  or  suction  air  chamber; 
steam  gauge,  5  inches  diameter;  water  gauge,  5  inches  diameter; 
relief  valve  of  large  capacity;  relief  valve  discharge  cone;  set 
of  brass  priming  pipes  and  special  priming  valves;  required 
number  of  2 j~ inch  Ludlow  hose  valves;  one  pint  sight-feed 
cylinder  lubricator;  one  pint  hand  oil  pump,  all  according  to 
the  specifications. 

On  account  of  the  larger  passageways,  brass  parts  and 
attachments  mentioned,  the  pump  necessarily  costs  more  than 
an  ordinary  fire  pump;  but  the  cost  by  the  gallon  discharged 
is  less,  since  the  underwriter  pump  can  deliver  a  greater  quantity 
of  water  in  the  same  length  of  time.  It  is  also  much  heavier 
and  stronger,  with  superior  workmanship,  and  better  protected 
from  rust  and  accident. 

Hose  valves  are  threaded  only  when  specially  ordered  and 
at  extra  cost.  If  these  valves  are  to  be  threaded  a  sample 
thread  must  be  supplied  by  the  purchaser,  as  there  is  no  estab- 
lished standard,  hose  threads  varying  widely  in  different  local- 
ities. 


Diameter 
of  Steam 
Cylinder, 
Inches. 

Diameter 
of  Water 
Plunger, 
Inches. 

Length  of 
Stroke, 
Inches. 

Normal 
Capacities 
Gals,  per 
Minute. 

Number 
of  Fire 
Streams. 

Steam 
Pipe, 
Inches. 

Exhaust 
Pipe, 
Inches, 

Suction 
Pipe, 
Inches. 

Dis- 
charge 
Pipe, 
Inches. 

14 

7l 

12 

500 

2 

3 

4 

8 

6 

16 

9 

12 

75° 

3 

3* 

4 

IO 

8 

18 

10 

12 

IOOO 

4 

4 

5 

12 

8 

2O 

12 

16 

1500 

6 

5 

6 

14 

IO 

Underwriter  fire  pumps  are  built  with  compound  steam  ends, 
and  power  underwriter  fire  pumps  are  arranged  to  run  with 
electric  motors. 

Pressure  Pumps  (Fig.  139)  are  those  which  are  intended 
for  hydraulic  pressure  installations  in  which  pressures  of  500 
pounds  and  over  are  used  to  operate  presses.  Although  these 


MODERN  RECIPROCATING  PUMPS  147 

pumps  are  sometimes  of  the  fly-wheel  tpye,  the  intermittent 
operation  of  the  pump  in  many  plants  makes  this  type  unsuit- 
able, hence  the  direct-acting  type  is  used  very  extensively. 
This  pump  is  marked  by  small  water  cylinders,  usually  of  the 
outside-packed  plunger  type.  The  valves  are  placed  in  valve 
boxes  bolted  to  the  water  end  on  account  of  the  complex 
casting  which  would  result  if  these  parts  were  integral  portions 
of  the  water  cylinders. 

The  Mine   Pump  (Fig.  140)  is  a  form  of  pressure  pump  as 
it  is  used  to  pump  water  from  great  depths.     The  water  end  is 


FIG.  140. — Compound  Outside-Packed  Mine  Pump. 

therefore  quite  similar  in  detail  to  the  pressure  pump.  There 
are  other  reasons  for  the  construction  of  the  water  ends  in 
small  parts.  The  action  of  the  acid  of  mine  waters  is  such 
that  parts  of  the  pump  are  rapidly  corroded,  hence  the  use  of  a 
number  of  small  parts  of  similar  form  makes  it  possible  to 
renew  the  part  affected  and  not  replace  a  whole  end,  as  would 
be  necessary  if  the  water  end  was  one  complex  casting.  More- 
over the  number  of  different  parts  is  reduced  to  a  minimum  by 
having  most  of  the  small  parts  similar  for  both  sides  of  the 
pump.  This  cuts  down  the  number  of  spare  parts  to  be  carried 
in  the  storeroom. 

The  table  below  gives  the  sizes  used  by  the  Worthington 


148 


PUMPING  MACHINERY 


Co.  for  their  Lehigh  pattern  of  pump  used  with  pressures  of 
300  pounds  per  square  inch  and  over. 


THE   WORTHINGTON    MINE    PUMP 
Lehigh  Pattern — For  300  Pounds. 

The  Worthington  mine  pump  (Lehigh  pattern)  is  specially 
designed  to  withstand  the  heavy  pressures  encountered  in 
deep  workings.  These  pumps  can  be  arranged  to  operate 
either  non-condensing  or  condensing,  and  are  also  made  with 
compound  and  triple-expansion  steam  cylinders,  where  a 
saving  of  fuel  is  desirable.  Pumps  of  this  design  of  larger 
size  or  for  heavier  service  can  be  furnished. 


1 

I 

1 

\ 

Minute. 

Size  of  Pipes  for  Short 
Lengths  to  be 
Increased  as  Length 
Increases. 

Approximate 
Space  Occupied. 
Feet  and  Inches. 

*0  !« 

*0  ,fl 

fc& 

CO 

* 

is 

l« 

1 

' 

^ 

gja 
c  >> 
.2o 

SB 

II 

Q 

M 

I 

•J1- 

33 

"o.S 

Is 

Gallons 

ai 
«s 

C/2 

3oi 
cd  5. 

•as 

W 

|4 

f 

H 

|£ 

jj 

Width. 

16 

6 

IO 

4.89 

43 

210 

•i 

3 

6 

5   • 

9     ii 

5      2 

i8i 

6 

10 

4-89 

43 

2IO 

3 

3i 

6 

5 

IO         O 

5     2 

16 

7i 

IO 

7.64 

43 

328 

2i 

3 

6 

5 

9     ii 

5     2 

i8i 

•7i 

10 

7.64 

43 

328 

3 

3i 

6 

1  5 

IO         O 

5     2 

*7  ' 

6 

IS 

7-35 

32 

235 

ai 

3i 

6 

5 

II       IO 

5     2 

20 

6 

IS 

7-35 

32 

235 

4 

5 

6 

5 

II     II 

5     2 

17 

7i 

IS 

11.48 

32 

367 

M 

3i 

6 

5 

II       IO 

5     2 

20 

7i 

IS 

ii  .48 

32 

367 

4 

5 

6 

5 

II     II 

5     2 

18 

7 

18 

12  .00 

28 

336 

3 

4 

6 

5 

22 

7 

18 

12.00 

28 

336 

4 

6 

6 

5 

18 

8 

18 

IS-67 

28 

438 

3 

4 

10 

8 

22 

8 

18 

IS-67 

28 

438 

4 

6 

IO 

8 

An  additional  charge  is  made  when  pumps  are  fitted  with  brass  plungers 
and  brass-bushed  glands. 

To  designate  the  sizes,  give  the  diameters  of  the  steam  cylinders  and 
water  plungers,  and  the  length  of  stroke. 

NOTE. — One  revolution  means  four  strokes,  counting  both  sides. 

When  lower  pressures  are  present  and  the  water  is  not 
corrosive  such  a  type  shown  in  Fig.  141  may  be  used.  The 


MODERN  RECIPROCATING  PUMPS 


149 


pump  is  of  simple  construction,  with  outside-packed  plungers. 
The  following  table  gives  sizes  used  for  this  type  of  pump. 

THE    WORTHINGTON    PACKED-PLUNGER   PUMP 
Scranton  Pattern — For  250  Pounds  Water  Pressure. 


6 

1 

g 

| 

c 

& 

Minute. 

Size  of  Pipes  for  Short 
Lengths  to  be 
Increased  as  Length 
Increases. 

Approximate 
Space  Occupied, 
Feet  and  Inches. 

11 

Is 

C/3 

"o 

4 

la 

1 

^ 

. 

oj  C 

6^ 

D  C 

6.3 

3. 

1! 

3  3 

"o.S 

1 

II 

is. 

•  ^4     PM 

C    OJ 

>  O. 

xj 

M 

.d 

§ 

9* 

|S 

cs 

IS 

•gs 

y'd, 

"«£ 

C 
flj 

12 

Q 

H 

_1 

o 

M 

(73 

« 

V) 

Q 

J 

^ 

14   . 

84 

10 

9-56 

54 

Sl6 

a| 

3 

8 

6 

9       8 

3      2 

16 

84 

10 

9-56 

54 

5l6- 

*4 

3 

8 

6 

9       9 

3    10 

18} 

8| 

10 

9-56 

54 

516 

3 

3^ 

8 

6 

9     10 

4     o 

16 

10} 

10 

J3-95 

54 

753 

*i 

3 

10 

8 

10       9 

3    10 

184 

loi 

10 

J3  -95 

54 

753 

3 

34 

10 

8 

10       9 

4     o 

18$. 

12 

10 

19.  16 

54 

1035 

4 

34 

12 

IO 

ii        i 

4     o 

20 

I  2 

IO 

19.16 

54 

1515 

H 

5 

12  • 

TO 

11          2 

4      2 

T7 

8* 

'5 

14.14 

40 

565 

4 

3^ 

8 

6 

10       5 

3    ii 

20 

84 

'5 

14.14 

40 

5*5 

a| 

5 

8 

6 

10       6 

4     2 

17 

10} 

15 

20.83 

40 

833 

4 

34 

10 

8 

TI          6 

3    ii 

20 

10} 

15 

20.83 

40 

833 

4 

5 

10 

8 

ii       8 

4      i  ^ 

20 

12 

15 

28.78 

40 

1151 

5 

5 

12 

10 

ii       9 

4     3 

A  sinking  pump  is  one  used  in  mines  for  dropping  into  a 
shaft  which  has  to  be  freed  of  water  collected  during  a  time 
of  disuse  or  which  has  come  in  when  a  subterranean  stream 
has  been  cut  into.  The  pump  is  subject  to  rough  usage  and 
must  be  applied  quickly  and  easily  to  the  timbers  of  the  shaft. 
Fig.  142  shows  one  form  of  sinker  as  built  by  the  Prescott 
Company.  This  pump  is  mounted  on  a  frame  which  can  be 
hung  from  the  mine  timbers.  The  working  parts  are  protected 
from  injury  from  falling  objects  by  the  projections  of  the  cyl- 
inder and  steam  chest.  The  chains  or  rods  connecting  the 
link  serve  to  support  the  pump  from  the  derrick.  The  water- 
valve  boxes  are  accessible  for  repairs  and  the  outside-packed 
plungers  make  it  possible  to  be  sure  that  there  is  no  leakage 
from  side  to  side.  When  electricity  is  applicable  the  electric- 


150 


PUMPING  MACHINERY 


sinking  pump  of  the  form  shown  in  Fig.  143  is  used.  This 
pump  is  driven  by  the  motor  through  a  double  reduction  gear 
and  the  plungers  are  of  the  form  mentioned  earlier.  The 
method  of  lowering,  the  protection  of  the  parts  from  mechanical 
injury  and  the  motor  from  water,  as  well  as  the  arrangement  of 
the  valves  for  quick  inspection  are  clearly  seen.  In  each  of 
these  sinking  pumps  the  suction  is  connected  to  the  bottom 
of  the  pump  and  the  discharge  may  be  taken  from  the  right 


FIG.  141. — Mine  Pump  Scranton  Pattern. 
(Size  16  and  25X14X15.) 

or  the  left.  These  pumps  are  so  built  that  they  may  run  if 
flooded  by  water,  as  such  a  condition  may  arise  at  any  time. 
The  table  on  page  151  gives  the  sizes  used  with  the  Knowles 
electric  duplex  sinking  pump. 


THE    KNOWLES    VERTICAL    DUPLEX    ELECTRIC    SINKING 

PUMPS 

Double- Acting  Outside  Center- Packed  Plunger  Pattern. 

This  pump  is  light,  compact,  efficient,  and  of  good  capacity, 
and  not  liable  to  damage  from  moisture  or  hard  usage.  The 
entire  motor  mechanism  is  inclosed  in  a  tight  casing,  but  every 


MODERN  RECIPROCATING  PUMPS 


151 


1 

c 

U 

Dimensions  over  all 

o 

c 

<8 

"3 

u 

• 

. 

_a 

g 

o 

in  Feet  and  Inches. 

<j    B 

*l 

1 

a 
33 

C/3 

is 

£-2 
.2 

L 

a 
cS! 

L 

o 

s 

1.1 

V  C 

o 

CJ  •*•* 

^  ^ 

c*S 

c  3 

OJ3 

rtX 

^    I., 

O  t^ 

^  ^ 

pi 

+J 

Ej3 

T* 

._<   0) 

"o.S 

J  « 

JJ 

•g  o 

Xo 

.2  g 

2  $ 

D1 

$1 

*c 

.<s£ 

2 

fefc 

SHH 
* 

1 

*  rl  ^ 

rtO 

PHp/ 

§ 

a 

<n 

Q 

00 

> 

M 

O 

O 

09 

^ 

ffi 

^ 

Q 

W 

3 

6 

400 

55 

•73 

40 

3 

2 

3  •  5  to  i 

IO 

3      o 

2      9 

8  8 

4 

6 

20O 

80 

1.30 

104 

4 

3 

2  .  7  to  i 

7i 

3      o 

2     9 

8  8 

4 

6 

500 

80 

1.30 

104 

4 

3 

3  .  7  to  i 

5 

20 

3      8 

3      6 

9  4 

o 

5 

6 

300 

75 

2  .04 

153 

•   5 

4 

3  .  7  to  i 

a 

20 

3      8 

4     2 

10    2 

5 

6 

4OO 

80 

2  .04 

163 

5 

4 

3  .  7  to  i 

*-*—  i 

0 

3° 

3      8 

4     2 

10  4 

6 

6 

200 

86 

2-93 

252 

6 

5 

3  .  7  to  i 

•d 

0) 
0) 

20 

3      9 

4     2 

10  4 

6 

6 

250 

86 

2.93 

252 

6 

5 

3  •  7  to  i 

g. 

25 

3      9 

4      2 

10  4 

6 

8 

400 

62 

3-91 

242 

6 

5 

3  •  5  to  i 

•g 

50 

4      6 

4   10 

II    2 

64 

8 

3OO 

70 

4-59 

321 

6 

5 

3  •  5  to  i 

w 
O 

5° 

4      6 

4   10 

112 

H 

7 

8 

300 

68 

5-32 

362 

8 

6 

3  •  5  to  i 

5° 

4      6 

4   10 

112 

8 

8 

300 

58 

6.96 

403 

8 

6 

3  •  5  to  i 

50 

4      6 

4   10 

112 

part  of  motor  and  pump  is  readily  accessible  for  examination 
or  repairs.  This  apparatus  will  stand  hard  usage  without  injury. 
The  main  frame  is  designed  especially  to  receive  the  specified 
type  and  size  of  motor.  As  the  plungers  and  piston  rods  are 
packed  from  the  outside,  any  leakage  is  readily  noticeable. 
Each  pump  is  furnished  with  a  discharge  air  chamber.  In 
sending  inquiries  or  orders,  always  state  as  fully  as  possible  the 
intended  service  and  requirements. 

Centrifugal  pumps  are  now  being  used  for  this  purpose 
at  times. 

The  Water-Works  Pump  will  be  considered  in  a  later 
chapter  in  more^  detail,  but  at  this  point  two  simple  forms  will 
be  described.  It  is  this  kind  of  pump  which  has  reached  the 
highest  form  of  development,  because  such  machines  are  used 
for  the  sole  purpose  of  raising  water  in  large  quantities  and 
continuously.  They  may  be  of  the  direct-acting  type  (Fig.  144) 
or  the  fly-wheel  type  (Fig.  145).  .  They  are  usually  compound  or 
triple  expansion  on  the  steam  end,  and  precautions  are  taken 
to  cut  down  all  losses.  The  pumps  are  made  with  great  refine- 
ment of  parts,  as  they  are  usually  handled  by  skilled  engineers. 


152  PUMPING   MACHINERY 

In  most  cases  there  is  a  possibility  of  adjusting  the  steam- 


FIG.  142. — Prescott  Steam  Sinking 
Pump. 


FIG.  143. — Electric  Sinking  Pump. 
(Size  6X6.) 


valve  gearing,  and  the  water  valves  may  be  examined  with 
ease.     These  pumps  vary  in  size  from  those  handling  50,000 


MODERN  RECIPROCATING  PUMPS  153 

gallons  per  twenty-four  hours  to   those  handling  30,000,000 


FIG.  144. — Direct-Acting  Water- Works  Pump. 
(Size  1 6  and  25X14^X18.) 


FIG.  145. — Water- Works  Pump. 

gallons  in  the  same  time.     Such  pumps  are  usually  rated  in 
this  manner — gallons  per  twenty- four  hours. 


154 


PUMPING   MACHINERY 


In  connection  with 
water  works  it  becomes 
necessary  to  lift  water  from 
artesian  or  deep  wells,  and 
for  this  work  special  deep- 
well  pumps  have  been 
designed.  These  consist 
usually  of  a  steam  cylinder 
mounted  at  the  top  of  the 
deep  well  (Fig.  146).  The 
piston  rod  of  such  a  pump 
is  usually  made  into  a 
plunger  passing  through  a 
stuffing  box  at  the  base  of 
the  pump  frame.  This 
plunger  is  then  connected 
to  a  bucket  piston  by 
wooden  rods  with  iron 
joints.  The  bucket  may 
be  several  hundred  feet 
below  the  surface.  This 
distance  is  fixed  by  the 
height  at  which  the  water 
stands  in  the  well  when  the 
pump  is  in  operation.  The 
foot  valve  is  placed  at  the 
end  of  the  pipe  line  which 
forms  the  pump  barrel. 
This  foot  valve  is  placed 
in  position  by  the  pump 
rods,  although  in  some 
cases  it  is  lowered  into 
position  by  other  means, 
its  weight  holding  it  after 
placing.  The  pump  bucket 
as  shown  by  the  figure  is 

FIG.  146. — Fairbanks-Morse  Deep- Well  Pump,   packed    by    means    of    CUp 


MODERN  RECIPROCATING  PUMPS  155 

leathers  and  the  foot  valve  and  head  valve  shown  are  of  the 
ball  type.  This  last  section  of  pipe  casing  seen  in  the  figure 
may  be  of  heavy  drawn  brass  tubing  to  make  a  proper  pump 
barrel.  In  arranging  the  plunger  at  the  top  of  the  wooden 
sucker  rods,  it  should  be  made  of  net  area  equal  to  one-half 
the  net  area  of  the  bucket  so  as  to  discharge  water  on  both 
strokes  of  the  pump. 

The  pump  frame  is  so  constructed  that  it  may  be  moved 
over  from  the  top  of  the  well  in  order  that  the  rods  or  piping 
can  be  taken  up.  This  is  done  by  hoisting  the  rods  or  casings 
until  one  section  is  above  ground,  then  clamping  the  line  by 
the  bottom  section  until  the  upper  section  is  unscrewed  and 
removed,  when  the  operation  is  repeated.  For  this  reason  it 
is  necessary  to  construct  the  pump  house  over  a  deep  well 
with  sufficient  head  room  to  permit  the  removal  of  casings 
or  rods;  this  is  usually  about  20  or  25  feet.  The  tables  below 
give  data  from  catalogue  of  the  Fairbanks-Morse  Co.  in  regard 
to  their  pumping  engines. 

THE   FAIRBANKS-MORSE   ARTESIAN- WELL   ENGINE 

This  engine  is  placed  directly  over  the  well,  and  the  piston 
rod  is  continued  to  the  required  depth  and  connected  to  the 
pump  piston.  The  steam  valve  is  perfectly  controlled,  and 
the  speed  of  the  engine  on  both  up  and  down  strokes  is  uniform. 
The  apparatus  may  be  run  at  a  high  speed  without  excessive 
shock  or  jar. 

These  engines  will  pump  from  the  deepest  wells,  forcing 
the  water  in  a  steady  stream  into  an  elevated  tank  or 
other  reservoir.  To  remove  the  pump  rods  and  pistons, 
the  bolts  which  connect  base  to  the  frame  are  loosened  and 
the  steam  cylinder  and  uprights  are  drawn  back  on  the 
base  by  a  screw.  The  upper  displacing  cylinder  discharges 
one-half  the  volume  pumped  on  the  down  stroke,  thus  tend- 
ing to  balance  the  machine  and  insure  a  smooth  and  easy 
action. 


156 


PUMPING  MACHIXERY 


SIZE  OF  ENGINE. 

SIZE  OF  PIPE. 

Diameter  of 
Steam  Cylinder. 

Length  of 
Stroke. 

Steam. 

Exhaust 

Inches. 

6 

1  8 

I 

li 

48     X20 

8 

24 

*l 

I* 

5l£X23 

10 

36 

*! 

2 

62     X25 

12 

36 

i*  • 

2 

62    X25 

14 

36 

2 

3 

73    X34i 

16 

36 

2 

3 

73    X34i 

BRASS   ARTESIAN-WELL    CYLINDER 


Inside 
Diameter. 

Length  of 
Stroke. 

Capacity  per 
Stroke  in 
Gallons. 

Outside 
Diameter  of 
Caps. 

Top  and 
Bottom  Con- 
necting Pipe  in 
Inches. 

2i 

18 

•31 

3& 

2^ 

2f 

18 

.46 

3i 

3 

3i 

18 

.64 

4ik 

3i 

3i 

18 

.86 

5* 

4 

1 

24 

.61 

3i 

3 

3i 

24 

.86 

43^ 

3* 

3i 

24 

.84 

5* 

4 

4\ 

\ 

24 

47 

si 

4) 

4! 

24 

.84 

6J 

5 

si 

24 

.68 

7i 

6 

3; 

36 

.29 

4i$ 

3i 

36 

•71 

si 

4 

4i 

36 

2.21 

5l 

4i 

4^ 

36 

2.76 

61 

5 

si 

36 

4-O2 

7i 

6 

TABLE  OF   WOOD   SUCKER   RODS 


Diameter  of  Rod  in  Inches, 
it 


Adapted  for  Working  Barrels  of  a  Diameter 
2!  to  41 
4l  to  sf 
5t  to  7! 


A  gas  engine  geared  to  a  deep-well  pump  is  shown  in  Fig. 
147,  in  which  the  same  points  are  to  be  noted  as  in  Fig.  146, 
while  Fig.  148  illustrates  the  method  of  using  a  horizontal 
cylinder  for  this  purpose,  as  proposed  by  Davidson. 

A  form  of  deep-well  power  pump  used  in  the  West  is  shown 
by  Fig.  149.  In  this,  the  Luitwieler  pump,  motion  is  given 


MODERN  RECIPROCATING  PUMPS  157 


FlG.  147. — Gas  Engine  Drive  for  Deep-Well  Pump. 


FIG.  148. — Davidson  Horizontal  Steam  Cylinder  for  Deep- Well  Pump. 


158 


PUMPING  MACHINERY 


to  the  spur  wheel  A  by  a  belt,  steam 
engine,  gas  engine,  or  electric  motor, 
and  this  drives  the  gear  B.  On  the 
gear  shaft  are  two  cams  arranged 
opposite  each  other  and  so  made  that 
they  have  a  lifting  motion  during  53 
per  cent  of  the  rotation,  while  the 
\  descending  motion  occupies  47  per 
J  cent  of  a  revolution.  These  cams 
/  drive  cross-heads  by  means  of  fric- 
J  tionless  rollers.  The  left-hand  cross- 
head  is  attached  at  its  lower  end  to 
a  hollow  pump  rod  and  the  right-hand 
one  at  its  upper  end  to  a  solid  rod 
passing  through  the  hollow  one.  The 
cross-heads  are  guided  to  prevent 
turning.  The  hollow  rod  passes 
through  a  stuffing  box  and  has  at  its 
upper  end  a  stuffing  box  for  the  solid 
rod.  These  rods  are  each  attached 
to  a  bucket  piston  as  shown  by  the 
figure.  The  outer  rod  is  made  of 
larger  bore  than  the  diameter  of  the 
solid  rod  so  that  there  is  no  friction, 
provided  the  rod  does  not  whip  and 
is  vertical.  When  the  lower  piston 
is  descending  the  upper  one  is  ascend- 
ing and  forcing  the  water  out,  while 
water  will  flow  in  between  the  two 
pistons  to  fill  the  space  formed  there. 
Before  the  upper  piston  reaches  the 
top  of  its  stroke  the  lower  piston 
starts 'to  move  upward  and  when  the 
upper  piston  begins  to  descend  the 
lower  piston  has  sufficient  speed  up- 

F.C.  i49.-Luitwieler  Deep-Well  Ward  tO  ke6P  the  COlumn  °f  Water  in 

Pump.  the  pipe  in  motion  at  the  same  ppeed. 


MODERN  RECIPROCATING  PUMPS  159 

The  water  lifted  by  the  lower  bucket  is  forced  through  the 
upper  bucket.  Before  the  lower  bucket  ceases  to  move  upward 
the  upper  one  has  completed  its  down  stroke  and  has  started 


FIG.  150. — Luitwieler  Quadruplex  Pump. 

back  again  ready  to  continue  the  upward  motion  of  the  water 
in  the  pipe  as  soon  as  the  lower  bucket  begins  to  diminish  its 
upward  rate  and  reverse  its  action. 

The   buckets,   rods,   and   cross-heads    balance    each    other 
quite  well,  so  that  there  is  a  fairly  uniform  torque  on  the  driving 


ICO  PUMPING  MACHINERY 

mechanism.  The  action  is  different  from  the  ordinary  deep- 
well  pump  and  its  differential  plunger  in  that  not  only  the 
stream  above  ground  is  moving  but  also  that  in  the  well  barrel. 
The  pump  lifts  water  on  the  suction  side  during  one-half  revolu- 
tion, as  in  the  case  of  the  ordinary  pump,  but  the  discharge  of 
this  through  the  well  barrel  is  distributed  over  the  whole  revo- 
lution. If  it  were  possible  to  put  the  upper  plunger  of  the  deep- 


FIG.  151. — Triplex  Pump. 
(Size  5X8.) 

well  pump  down  near  the  working  barrel  the  same  action  would 
occur  in  that  pump.  The  arrangement  of  cams  may  be  such 
that  the  motion  of  the  water  is  uniform,  the  accelerating  period 
of  the  motion  of  one  occurring  during  the  time  in  which  the 
other  is  changing  its  motion. 

The  same  arrangement  of  cams  may  be  applied  to  duplex 
or  quadruplex  pumps,  where  the  motion  given  by  the  ordinary 
crank  is  not  sufficiently  uniform.  Such  an  arrangement  is 


MODERN  RECIPROCATING  PUMPS 


161 


shown  in  Fig.  150.     In  this  case  the  effort  is  made  to  sub- 
merge the  pump  barrels  to  cut  down  all  suction  lift. 

This  leads  to  a  large  class  of  pumps  known  as  power  pumps. 
Power  pumps  are  those  in  which  the  pump  proper  is  driven 
through  gears  or  belts.  These  pumps  are  of  various  forms, 
some  horizontal,  some  vertical,  some  with  plungers,  some  with 


FIG.  152. — Horizontal  Duplex  Power  Pump. 
(Size  6X12.) 

pistons,  some  with  single  cylinders,  some  with  double  cylinders, 
and  some  triplex.  The  name  triplex  usually  designates  a 
vertical,  three-cylinder  pump.  In  many  cases  it  is  of  the 
plunger  type.  Fig.  151  illustrates  this  form.  The  large  valve 
case  with  a  simple  arrangement  for  examining  the  valves,  the 
inlet  and  outlet  openings,  the  short  connections  to  the  cyl- 
inder, the  method  of  supporting  the  side  thrust,  the  method  of 


162 


PUMPING   MACHINERY 


packing,  and  the  arrangement  of  cranks  and  gears  are  clearly 
shown.     The  sizes  of  such  pumps  are  given  in  the  table  below. 


DEANE  TRIPLEX  VERTICAL  SINGLE-ACTING  POWER  PUMP 


Size. 

I 

4 

Capacity. 

Pipe 
bizes. 

Tight  and 
Loose  Pulleys. 

j 

Approximate 
Dimen.  in  Feet 
and  Inches. 

c 

<r.  U 

'£ 

0) 

VM 

eg 

II 

§S 

«  . 

&>„ 

. 

O 

il 

5 

Length  c 

Stroke 

Gallons  i 
Revolt 

Revoluti 
per  Mir 

Il 

O 

Suction. 

Dischar^ 

Diamete 

d 

1 

"3 
PQ 

Ratio  of 

! 

Width. 

Height. 

3 

8 

290 

-73 

45 

33 

3 

24 

3° 

54 

D 

4.7-1 

3    10 

3      9 

5  7 

4 

8 

290 

1.30 

45 

58 

3 

24 

3° 

54 

D 

5-o-i 

4      8 

4    ii 

6  5 

4 

8 

270 

1.30 

45 

58 

3 

24 

30 

54 

D 

4.7-1 

3    10 

3      9 

5  7 

44 

8 

290 

1-65 

45 

74 

4 

3 

3° 

D 

5-o-i 

4      8 

4    ii 

6  5 

5 

8 

254 

2.04 

45 

91 

4 

3 

30 

74 

D 

5-0-1 

4      8 

4    ii 

6  5 

5 

8 

180 

2  .04 

45 

91 

4 

3 

3° 

6* 

D 

4-7-1 

3    10 

3      9 

5  7 

54 

8 

209 

2.46 

45 

no 

4 

3 

3° 

74 

D 

5-o-i 

4      8 

4    ii 

6  5 

54 

8 

150 

2  .46 

45 

I  10 

4 

3 

3° 

64- 

D 

4-7-1 

3    10 

3      9 

5  7 

6 

8 

200 

2-93 

4b 

132 

4 

3 

30 

74 

D 

5-o-i 

4     8 

4    ii 

6  5 

6 

8 

125 

2-93 

45 

132 

4 

3 

3° 

D 

4-7-i 

3    10 

3      9 

5  7 

64 

8 

150 

3-44 

45 

154 

I5 

24 

36 

74 

D 

5-o-i 

4     8 

4   ii 

6  5 

64 

8 

107 

3-44 

45 

24 

3° 

64 

D 

4-7-1 

3    10 

4      i 

5  7 

7 

8 

150 

4.00 

45 

180 

2  5 

24 

36 

84 

D 

5-o-i 

4      84   ii 

6  5 

7 

8 

92 

4.00 

45 

180 

25 

24 

30 

74 

D 

4-7-1 

3    10 

4      i 

5  7 

8 

8 

IOO 

5-22 

45 

235 

26 

*  5 

36 

84 

D 

5-o-i 

4     8 

4   ii 

6  5 

1  D  indicates  double  belt.  2  Flanged  connections. 

Above  pumps  may  be  safely  run  at  somewhat  greater  speed  than  listed. 


The  pumps  listed  above  have  outside-packed  plungers  which 
work  through  deep-stud  gland  stuffing  boxes.  All  sizes  have 
cross-heads  for  guiding  the  travel  of  the  plungers,  and  approved 
means  provided  for  taking  up  wear.  The  bearings  are  of 
ample  diameter  and  length.  The  gearing  is  a  special  mixture 
of  steel  and  iron,  and  the  teeth  are  machine  cut.  The  con- 
necting rods  are  provided  with  adjustable  boxes  at  both  ends, 
the  adjustment  being  made  by  means  of  a  wedge  and  screw. 

Pumps  of  this  type  are  used  for  general  service  and  tank 
service;  as  boiler  feeders,  elevator  pumps,  and  water  works. 

This  type  of  pump  is  built  with  3-inch,  6-inch,  8-inch,  and 


MODERN  RECIPROCATING  PUMPS 


163 


12-inch  stroke,  each  class  having  the  same  general  character- 
istics as  described  above. 

Fig.  152  shows  a  horizontal  duplex  power  pump  with 
pistons.  There  are  many  different  forms  of  this  type  of  pump, 
but  these  two  will  illustrate  'the  ordinary  styles. 

Fig.  153  illustrates  the  power  head  for  a  deep-well  pump, 


FIG.    153. — Power  Head  for  Deep- Well  Pump. 

giving  the  method  of  operating  such  a  pump  with  belting. 
An  electric  motor  could  be  attached  by  a  coupling  or  belt 
to  the  shaft  carrying  the  belt  pulley.  The  figure  shows  the 
method  of  removing  the  top  gearing  from  the  well  when  it  is 
necessary  to  remove  rods  or  casing. 

A  new  form  of  power  pump  usually  driven  direct  from 
an  electric  motor  without  the  use  of  gears  to  cut  down  the 


164 


PUMPING  MACHINERY 


speed  is  the  express  pump.  Such  a  pump,  Fig.  154,  is  arranged 
to  run  at  high  rotative  speeds  of  about  200  to  300  R.P.M.  The 
Riedler  pump  of  the  Allis-Chalmers  Company  is  arranged  to 
permit  this  high  speed  by  having  a  clear  direct  passage  for  the 
water  through  large  valves  and  by  making  the  discharge  valve 
positive  in  its  closing.  The  water  enters  through  the  annular 
space  AB  from  the  vacuum  air  chamber  C.  It  is  then  dis- 
charged into  the  air  chamber  K  through  the  valve  E.  This 


FIG.  154. — Allis-Chalmers'  Duplex  Riedler  Express  Pump 

valve  is  closed  positively  by  means  of  an  eccentric  on  the 
shaft.  The  plunger  GI  is  of  one-half  the  area  of  G  so  that 
water  is  passed  back  and  forth  through  F,  giving  a  discharge 
on  each  stroke,  although  there  is  only  one  suction  stroke  to 
each  revolution. 

The  ring  D,  carried  from  the  back  of  the  plunger,  closes 
the  annular  valve  when  the  plunger  reaches  the  end  of  the 
suction  stroke,  thus  cutting  down  the  chance  of  slip  from  the 
slow  action  of  the  valve. 


MODERN  RECIPROCATING  PUMPS  165 

The  suction  air  chamber  C  insures  a  supply  of  water  at  all 
times. 

The  express  pumps  of  Riedler  were  among  the  earliest  of 
this  type,  but  in  later  years  there  have  been  many  new  forms 
introduced.  In  some  cases  the  valves  have  been  spring  con- 
trolled, eliminating  the  necessity  of  the  valve  gearing. 

The  simplex  pump,  the  water- works  pump,  the  centrifugal 
and  air-lift  pumps,  air  pumps  for  condensers,  and  other  special 
pumps  will  be  considered  in  later  chapters. 


CHAPTER  IV 
SIMPLEX  PUMPS 

THE  pump  containing  one  steam  cylinder  and  one  water 
cylinder  is  known  as  a  simplex  pump  to  distinguish  it  from 
the  duplex  pump,  so  largely  described  in  the  previous  chapter. 
This  type  was  the  original  form  used  by  Worthington,  and  it 
has  been  worked  on  by  many  inventors.  It  is  the  purpose 
of  this  chapter  to  describe  a  number  of  the  well-known  types 
which  will  show  the  general  manner  of  solving  the  problem 
of  getting  a  definite  reversal  of  the  pump  at  the  end  of  the 
stroke  no  matter  what  the  speed  of  the  pump  is.  This  has 
been  the  aim  of  all  simplex-pump  designers.  It  is  a  simple 
matter  to  reverse  the  action  when  the  pump  is  moving  rapidly, 
but  for  slow  action  there  must  be  some  auxiliary  apparatus, 
as  the  spring  of  the  first  Worthington  pump,  Fig.  55. 

The  simplex  pump  is  undoubtedly  more  complex  than  the 
duplex  pump,  but  with  it  there  is  little  danger  of  short  stroking. 
In  the  duplex  pump  the  valves  must  be  carefully  adjusted,  as 
the  motion  of  one  piston  controls  the  action  of  the  valve  of 
the  other  side.  The  simplex  pump  in  many  cases  may  be  run 
at  higher  speeds  than  are  common  with  duplex  pumps.  The 
loss  due  to  the'  clearance  space  is  a  total  loss  in  direct-acting 
pumps,  as  there  is  no  expansion,  and  for  this  reason  the  clear- 
ance should  be  made  as  small  as  possible  and  no  short  stroking 
should  be  allowed. 

The  Cameron  Pump  (Fig.  155)  is  one  of  the  simpler  forms 
of  these  pumps.  As  shown  in  the  figure  the  B- valve  G  is  ad- 
mitting steam  to  the  right-hand  side  of  cylinder  A  and  the 
piston  is  being  driven  to  the  left.  When  the  piston  C  over- 
runs the  steam  passage  at  its  left  end  the  exhaust  steam  is 
cushioned  partially  as  it  can  only  pass  through  a  groove  at  the 

166 


SIMPLEX  PUMPS 


167 


top  of  the  cylinder  to  the  passage.  The  piston  C  strikes  the  pin 
on  the  cylindrical  valve  7,  allowing  the  steam  to  the  left  of 
the  piston  F  to  escape  through  E  into  a  passage  leading  to  the 
exhaust.  The  steam  which  has  leaked-  into  the  space  at 
the  right  of  the  auxiliary  piston  F,  through  the  small  hole  in  the 
center  of  the  head,  drives  this  to  the  left,  moving  with  it 
the  valve  G,  and  reverses  the  pump.  When  the  piston  C  moves 
to  the  right,  live  steam  entering  behind  /  through  the  passage  K 


FIG.  155. — Cameron  Pump. 

forces  the  valve  /  over,  closing  the  passage  to  the  left  of  the 
cylinder  in  which  F  travels  while  the  steam  which  leaks  through 
the  small  hole  in  the  end  of  F  builds  up  the  pressure  for  the 
next  reversal,  when  the  right-hand  valve  /  is  opened.  The 
auxiliary  piston  F  controls  the  valve  G.  It  is  made  of  two 
hollow  pistons  with  holes  in  the  ends.  The  steam  which  leaks 
through  these  openings  represents  the  cost  of  operating  the 
valve.  This  may  be  an  appreciable  amount,  as  the  steam 
may  fill  the  space  at  the  left-hand  end  in  the  figure  before 


168  PUMPING  MACHINERY 

the  exhaust  occurs.  With  a  small  hole,  however,  there  need 
not  be  much  steam  used,  as  it  is  only  the  steam  on  the  smaller 
end  which  is  needed  to  reverse  the  valve.  The  handle  H  pro- 
jecting from  the  side  of  the  steam  chest  serves  to  move  the 
valve  back  and  forth  in  starting. 

The  figure  shows  the  arrangement  of  the  suction  and  dis- 
charge valves.  These  valves  are  controlled  by  springs,  and 
by  taking  out  a  common  spindle  the  valves  on  one  side  may  be 
removed.  The  clear  passage  for  the  suction  is  seen  and  the 
short  length  of  this  pump  in  comparison  with  its  stroke  is 
important  where  space  is  necessary. 

The  table  on  p.  169  gives  the  sizes  of  this  pump  used  for 
boiler  feeding  as  arranged  by  the  makers. 


FIG.  156 — Cameron  Pump. 

CAMERON   PISTON    PUMP— REGULAR   BOILER-FEED 
PATTERN 

The  main  difficulty  met  with  in  fixing  on  the  proper  size 
of  pump  to  recommend  is  that  the  horse  power  of  the  boiler 
for  which  the  pump  is  required  is  about  all  the  information 
furnished.  The  expression  "  horse  power,"  as  applied  to 
boilers,  is  a  very  indefinite  term;  what  should  be  given,  if 


SIMPLEX  PUMPS 


169 


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170 


PUMPING    MACHINERY 


possible,  is  the  quantity  of  feed  water  required,  and  a  pump 
which  will  supply  this  quantity  at  about  one-half  its  rated 
capacity  at  ordinary  speed  will  be  right  for  cold  water,  and  say 
one- third  speed  for  hot  water. 

In  feeding  hot  water  the  pump  should  be  placed  below  the 
source  of  supply.  The  first  six  sizes  are  furnisheu  wiin  hand- 
lever  attachment  when  so  desired. 

An  outside-packed  type  of  the  Cameron  pump  is  shown 


FIG.  157. — Marsh  Steam  Pump. 
(8X5X10,  8X5X12,  10X6X12.) 

in  Fig.  156.  The  view  not  only  shows  the  type  of  water  end, 
but  also  shows  the  reversing  valve  cylinders  on  the  steam 
cylinder  heads  and  the  reversing  handle. 

The  Marsh  Pump  (Fig.  157)  is  another  one  using  a  steam- 
thrown  valve.  In  the  picture  shown  the  piston  travels  to- 
the  right  while  the  steam  enters  the  cylinder  through  the  annu- 
lar space  around  the  spindle  valve  and  passes  into  the  steam 


SIMPLEX  PUMPS 


171 


passage.  Should  the  water  resistance  on  the  water  piston 
decrease  and  the  steam  piston  respond  suddenly  to  this,  the 
pressure  in  the  cylinder  would  fall  and  with  it  the  pressure 
on  the  left  of  the  disc  on  the  left  end  of  the  spindle-slide  valve. 
The  pressure  on  the  right-hand  side  of  this  disc  would  then 
move  the  spindle,  cutting  down  the  steam  supply  and  throttling 
the  exhaust,  thus  dampening  the  action  of  the  steam  piston. 


FIG.  158. — Marsh  Pump. 

When  the  piston  reaches  the  end  of  its  stroke  it  overrides 
the  small  port  leading  to  the  right-hand  end  of  the  valve  piston, 
and  the  live  steam  which  is  introduced  into  the  central  space 
between  the  two  parts  of  the  steam  piston,  then  passes  behind 
the  right-hand  side  of  the  right  disc  on  the  valve  spindle, 
and  forces  it  to  the  left.  This  admits  steam  to  the  right-hand 
end  of  the  steam  cylinder  and  opens  the  left-hand  end  to  the 


172 


PUMPING  MACHINERY 


exhaust.  It  is  clear  from  the  figure  that  the  two  faces  of  the 
right-hand  disc  are  exposed  to  exhaust  pressure  and  the  left- 
hand  disc  has  live  steam  on  each  side;  the  valve  is  thus  balanced 
until  the  central  part  of  the  main  piston  covers  the  auxiliary 
passage,  admitting  live  steam  on  one  side  and  thus  destroying 
the  balance  in  such  a  manner  that  the  valve  shifts  readily. 
The  dotted  lines  show  the  passages  carrying  steam  to  the  cen- 
tral part  of  the  piston  and  to  the  valve  case. 

The  balanced  feature  of  the  valve  means  that  the  valve  is 
adjusted  for  speed  so  that  the  pump  will  not  race. 

This  company  sometimes  builds  the  steam  piston  with  the 
central  space  as  wide  as  the  stroke  So  that  steam  may  be  intro- 
duced into  this  space  by  a  connecting  passage  in  the  cylinder 
casting,  saving  'the  packed  rod.  This  method  means  a  much 
longer  steam  cylinder  and  also  considerable  loss  from  steam 
condensation. 

The  table  below, gives  the  sizes  of  these  pumps: 

MARvSH    STANDARD    BOILER    FEED-PUMPS 
FOR  HIGH  PRESSURE.     REGULAR  LINE 


1 

1 

<L> 

0) 

o 

1 

If 

So 

C| 

2^ 
fc?° 

u 

1 

2  ^ 
o   ° 

W 

15  a 

ii 

11 

|| 

O  0) 

•«£ 

U  Cj 

§a 
_t/j 

A 

I 

o 

• 

^ 

« 

O 

0 

O 

09 

w 

CO 

q 

fe 

K 

^ 

7 

4l 

8 

•55 

66 

4,000 

I 

it 

3i 

«l 

15X40 

400 

650 

7i 

4i 

IO 

.68 

75 

4,5°° 

I 

Jl 

3i 

22 

18X48  450 

900 

8 

5 

IO 

•85 

85 

5,100 

I 

ij 

2  i 

18X48  500 

1000 

8 

5 

12 

i  .02 

IOO 

6,000 

I 

i£ 

4 

3 

18X52 

600 

1150 

10 

6 

12 

i  .46 

146 

8,75° 

1^ 

ii 

4 

3 

18X52 

1000 

1300 

12 

7l 

12 

2.14 

216 

13,000 

!| 

2 

5 

4 

19X56  1500 

2500 

12 

8 

12 

2.61 

260 

15,600 

l£ 

2 

5 

4 

19X56 

1650 

1500 

14 

8J 

12 

2-94 

295 

17,700 

2 

2^ 

6 

5 

22X62 

2OOO 

2400 

16 

10 

16 

5-44 

408 

25,000 

3 

3i 

8 

7 

24X76 

25OO 

4800 

16 

9* 

16 

4-65 

350 

21,000 

3 

3* 

8 

7 

24X76 

2200 

4800 

16 

10 

20 

6.80 

450 

27,000 

3 

3l 

8 

7 

24X76 

270O 

5000 

16 

9} 

20 

5.81 

400 

24,000 

3 

3i 

8 

7 

24X76 

24OO 

5000 

20 

12 

2O 

9.78 

588 

35,000 

3i 

4 

8 

7 

26X93 

35°° 

6500 

Capacity,  Capacity  rating  above  is  intended  to  represent  maximum 
service  recommended.  It  is  always  advisable  to  run  pumps  at  moderate 
speed,  consequently  select  sizes  large  enough  to  meet  ordinary  require- 
ments easily.  Reserve  power  in  a  pump  is  quite  as  important  as  in  9. 
boiler  or  engine. 


SIMPLEX   PUMPS 


173 


The  small  tappets  at  each  end  of  the  valve  casing  are  used 
to  start  the  pump  in  case  it'  does  not  operate.  Fig.  158  gives 
an  outside  view  of  this  pump  while  the  method  of  attaching  the 
suction  is  seen  in  the  two  figures.  Steam  may  be  exhausted 
into  the  base  where  the  suction  enters  by  turning  the  handle 
shown  in  Fig.  158.  The  exhaust  steam  heats  the  feed  water 
and  is  thus  saved  for  useful  application. 

One  of  the  oldest  of  the  simplex  pumps  is  the  Knowles 
pump.  The  Warren  steam  pump  is  quite  similar  to  it,  and 
its  action  will  be  described.  The  operation  of  this  pump  is 


FIG.  159. — Knowles  Pump. 

shown  in  Fig.  159.  When  the  steam  piston  reaches  the  left 
end  of  its  stroke  the  roller  K  strikes  the  lever  R,  forcing  up  the 
rod  L,  which  is  attached  to  an  arm  D,  projecting  from  the  rod 
of  the  auxiliary  piston  A.  This  rotates  A  through  a  small 
angle  and  brings  a  groove  in  A  extending  to  the  right  end  over 
a  small  steam  passage  while  a  groove  on  the  left  end  of  A  is 
brought  over  an  exhaust  passage.  This  forces  the  auxiliary 
piston  to  the  left,  moving  the  B- valve  into  the  position  shown, 
reversing  the  pump.  At  the  right-hand  end  of  the  stroke  the 
roller  K  strikes  the  other  end  of  R,  pulling  the  rod  L  downward. 
This  rotation  brings  the  groove  on  the  right  end  of  A  over  an 
exhaust  passage,  while  that  on  the  left  comes  opposite  a  steam 


174 


PUMPING  MACHINERY 


passage  and  the  auxiliary  piston  is  driven  to  the  right.     Should 
the  rotation  fail  to  reverse  the  pump,  the  rod  I  would  hit  the 


FIG.  160. — Knowles  Boiler  Feed  Pump. 
(Size  10X6X12.) 


FIG.  161. — Warren  Pump. 
(Size  7^X7^X10.) 

tappets  on  the  rod  from  the  auxiliary  piston  and  drive  the 
valve  to  its  opposite  position. 


SIMPLEX   PUMPS 


175 


The  roller  may  be  raised  and  lowered.  This  with  the  adjust- 
ment in  the  connecting  rod  L  makes  it  possible  to  adjust  the 
point  of  reversal. 

The  Knowles  heavy  pattern  boiler-feed  pump  is  shown  in 
Fig.  160,  while  Fig.  161  shows  a  Warren  light-service  pump. 

The  table  below  gives  sizes  of  the  Knowles  pump. 

THE   KNOWLES   HORIZONTAL   BOILER-FEED    OR   PRESSURE 

PUMPS 

Piston  Pattern — For  Hot  or  Cold  Water  or  Other  Liquids 

All  parts  of  this  pump  are  interchangeable,  and  can  therefore 
be  readily  duplicated  in  case  of  accidental  breakage  or  unusual 


wear. 


REGULAR    PATTERN.       FOR    125    POUNDS    WATER    PRESSURE. 


Capacity  per 

<3 

<5 

v> 

Minute  at  Max- 

Jv  " 

B 

•0 
a 

B 

o 

L 

<u 

imum  Speed. 

£ 

•i 

Pn 

•  <u 

ex 

.^H 

I 

£ 

sJ 

1 
B 

o£ 
|| 

ater  Cy 
Inches. 

fl 

i 

a<u 

3-g 
ij 

J 

a 

n 

§ 

^  ^ 

B-g 

PH     . 

C/3 

§^ 

fej 

Ij* 

fj 

11 

z 

09 

^ 

co 

o 

co 

0 

w 

w 

Q 

--H      W 

000 

»i 

i* 

3 

.023 

15° 

3i 

i 

I 

i 

f 

i8X    4 

00 

3 

if 

3 

.031 

i 

1 

| 

i8X    5 

0 

3i 

2 

4 

.05 

150 

72 

i 

1 

Ji 

I 

31X6 

I 

3i 

2i 

4 

.07 

150 

10^ 

i 

1 

ij 

I 

29X    7 

2 

4 

•| 

5 

.  II 

150 

i6i 

2 

1 

x¥ 

I 

32X    7 

3 

5 

3* 

7 

•25 

I25 

31 

I 

I 

2 

J2 

44X10 

4 

5i 

3l 

7 

•33 

I25 

41 

i 

I 

2 

1^ 

44X10 

4i 

7 

4 

7 

•38 

125 

47 

i 

tj 

22 

2 

45Xio 

6 

7i 

5 

10 

•85 

IOO 

85 

i- 

^i 

3 

2i 

56X16 

6i 

8 

5 

12 

I  .02 

IOO 

102 

i 

ri 

4 

4 

64X16 

7 

10 

6 

12 

1.47 

IOO 

147 

*i 

!i 

4 

4 

68X19 

8 

12 

7 

12 

2  .  OO 

IOO 

2OO 

2 

22 

5 

5 

68X19 

9 

14 

8 

12 

2.6l 

IOO 

26l 

2 

2i 

5 

5 

68X20 

10 

16 

10 

16 

5-44 

75 

408 

2i 

3 

6 

6 

81X21 



16 

10 

18 

6.12 

70 

428 

2i 

3 

8 

6 

96X22 

18 

IO 

18 

6.12 

70 

428 

22 

3 

8 

6 

96X24 

The  Nos.  o,  i,  2,  3,  4,  and  4$  pumps  are  provided  with  hand-power 
attachments.  By  this  means  the  pump  can  be  used,  when  steam  s 
down,  for  filling  .boilers  after  "blowing  off/'  washing  decks,  fire  purposes, 
etc.  The  hand  lever  can  be  easily  removed  by  simply  lifting  it  from 
the  pump. 

Nos.  ooo,  oo,  and  o  pumps  may  be  built  as  "reinforced  pattern," 
designed  for  400  pounds  pressure. 


176 


PUMPING   MACHINERY 


These  pumps  have  large  direct  water  passages  and  full 
valve  areas,  which  not  only  reduce  water  friction  to  a  minimum, 
but  enable  them  to  be  run  at  a  speed  that  makes  them  efficient 
fire  pumps. 

HEAVY   PATTERN.        FOR    250   POUNDS    WATER   PRESSURE 


I* 

| 

e/3 

Capacity  per 
Minuteat  Max- 

ti 

13 

c 

^5 

imum  Speed. 

a 

9 

a 

'& 

^ 

*>.    . 

CJ 

C 

s.  . 

0. 

£ 

a. 

o£ 

Steam  C 
Inches 

Water  C 
Inches 

CO 

Gallons 
Stroke 

Strokes. 

Gallons. 

11 

CO 

Exhaust 
Inches 

Suction 
Inches 

Delivery 
Inches 

co'S 

I"8 

7 

41 

10 

.69 

IOO 

69 

I 

zi 

3 

2.1 

60X14 

7$ 

5 

IO 

•8.S. 

IOO 

85 

I 

ij 

3 

2* 

60X14 

8 

5 

12 

I  .02 

IOO 

102 

I 

tl 

4 

4 

66X14 

10 

6 

12 

I  .47 

IOO 

*47 

1^ 

Ji 

4 

4 

69X15 

12 

7 

12 

2  .OO 

IOO 

200 

2 

2i 

5 

5 

72X20 

14 

8 

12 

2.6l 

IOO 

26l 

2 

2 

5 

5 

72X20 

Twice  the  above  capacities  can  be  had  in  emergencies ;  but  for  contin- 
uous work,  such  as  boiler  feeding,  about  half  the  speed  stated  is  advised. 

The  Blake  pump  (Fig.  162)  is  operated  in  a  different  manner. 
When  the  tappet  A  is  struck  as  the  piston  moves  to  the  left 
the  casting  B  containing  the  cavities  C,  Z),  and  E  is  moved 
to  the  left.  This  motion  causes  the  projection  F  on  the  casting 
B  to  cover  the  passage  carrying  steam  to  the  right  of  the  aux- 
iliary piston  G  while  the  projection  H  uncovers  the  passage 
leading  to  the  left  end  of  G.  The  projection  1  on  the  other 
side  of  B  contains  two  cavities  so  that  this  motion  to  the  left 
covers  the  passage  leading  to  the  left  end  of  the  auxiliary 
piston,  while  that  leading  to  the  right-hand  end  is  connected 
through  one  of  the  cavities  under  /  to  the  exhaust.  The  steam 
then  entering  the  left  of  G  when  the  cavity  at  the  right  is  con- 
nected to  the  exhaust  causes  the  piston  to  travel  to  the  right 
and  with  it  the  main  valve  K,  thus  reversing  the  pump.  The 
action  is  then  repeated  in  the  reverse  direction  when  the  arm 
strikes  the  tappet  M. 

Should  the  auxiliary  piston  fail  to  operate,  the  movement 
of  the  casting  B  is  such  that  the  left-hand  passage  would  be 
moved  to  the  left  so  far  that  steam  would  be  admitted  to  the 


SIMPLEX   PUMPS  177 

left-hand  end  of  the  main  cylinder,  while  exhaust  would  occur 


(~ 

p 

_____  •> 

3           1 

't%%%%%%^^            ^ 

r? 

tn 

rtrib 

L 

FIG.  162. — Blake  Steam  Pump. 


FIG.  163. — Deane  Pump. 

from  the  right.     Such  action  might  cease  if  the  pump  were 
running  slowly. 


178 


PUMPING  MACHINERY 


The  passages  operating  the  auxiliary  valve  are  made  in 
the  cylinder  casting,  and  although  small,  they  carry  sufficient 
steam  to  operate  the  auxiliary  piston.  The  casting  B  is  really 
the  valve  operating  the  auxiliary  piston.  It  contains  passages 
leading  to  the  main  passage  because  it  is  necessary  to  connect 
these  with  the  main  valve.  The  same  action  could  be  obtained 
if  the  valve  portion  of  this  was  made  like  a  frame  surrounding 


FIG.  164. — Blake  Pump. 


seat. 


the  main  valve  while  the  main  valve  moved  on  the  lowe1 
Such  an  arrangement  is  used  on  the  Deane  pump. 

The  Deane  pump  contains  a  frame  B  surrounding  the  main 
valve  /.  The  tappet  A  (Fig.  163)  is  struck  by  the  sleeve 
and  moves  a  frame  B  so  that  the  right-hand  end  of  the  auxiliary 
cylinder  is  connected  to  the  exhaust,  while  the  left  end  is  con- 
nected to  the  steam  supply.  In  this  way  the  auxiliary  piston 
is  driven  to  the  right  and  with  it  the  main  valve  is  carried  over, 
reversing  the  pump.  Should  the  auxiliary  piston  cease  to  act, 


SIMPLEX   PUMPS  179 

the  continued  motion  of  the  pump  would  pull  the  main  valve 
to  the  right  by  means  of  the  reverse  lever  M  and  thus  reverse 
the  pump.  The  frame  B  is  really  a  valve  operating  the  aux- 
iliary piston.  One  end  of  this  controls  the  admission  and 
exhaust  of  steam,  from  one  end  corresponding  to  one  cylinder 
end,  while  the  other  of  the  frame  controls  the  other  end. 
These  pumps  are  accurate  in  their  action.  Fig.  164  shows 


FIG.  165. — Deane  Pressure  Pump. 
(Size  i8X2jXi8.) 

the  exterior  view  of  the  Blake  pump  and  Fig.  165  shows  a 
long-stroke  pressure  type  of  the  Deane  pump,  to  which  the 
table  on  p.  180  refers. 

THE    DEANE    SINGLE    OUTSIDE-PACKED    DOUBLE-PLUNGER 
PUMP— POT-VALVE    PATTERN 

For  pressures  between  300  and  3000  Ibs.  the  style  of 
pump  shown  in  Fig.  165  is  recommended.  As  will  be  seen,  the 
arrangement  of  the  plungers  and  cylinders  is  similar  to  the 
regular  outside  plunger  type,  with  the  difference,  however, 
that  the  valves  are  placed  in  pots  above  the  cylinder.  These 
valves  are  specially  designed  for  the  service;  are  each  in  separate 
compartments,  and  may  be  readily  inspected  by  the  removal 
of  covers.  The  cylinders  are  made  of  steeline,  a  special  hydraulic 
cylinder  mixture,  for  all  pressures  up  to  and  including  1250 
Ibs.  For  higher  pressures  water  ends  are  made  of  open-hearth 
steel  castings.  These  pumps  are  adapted  for  hydraulic  cranes, 
presses,  punches,  shears,  riveting  machines,  etc.  All  parts  are 


180 


PUMPING  MACHINERY 


quickly  accessible,  devoid  of  complications,  and  heavy  in  con- 
struction. These  pumps  are  also  built  with  duplex  compound 
and  triple-expansion  steam  ends.  Water  ends  of  pumps  listed 
below  are  of  steeline  good  for  a  working  pressure  of  1250  Ibs. 


SIZE. 

CAPACITY. 

PIPE  SIZES. 

Diam- 
eter of 
Steam 
Cyl. 

Diam- 
eter of 
Plun- 
ger. 

Length 
of 
Stroke. 

Gallons 
per 
Stroke. 

Strokes 
per 
Minute  of 
One 
Plunger. 

Gallons  per 
Minute  of  Both 
Plungers. 

Diam- 
eter of 
Steam 
Pipe. 

Diam- 
eter of 
Exh'st 
Pipe. 

Diam- 
eter of 
Suc- 
tion 
Pipe. 

Diam- 
eter of 
Dis- 
charge 
Pipe. 

IO 

a* 

12 

.  21 

40  to  80 

8.  5  to     17 

I* 

2 

2 

li 

12 

a* 

12 

.  21 

40  to  80 

8  .  5  to    17 

2 

ai 

2 

li 

14 

H 

12 

.  21 

40  to  80 

8  .  5  to     17 

2 

a| 

2 

li 

16 

a* 

18 

•31 

40  to  80 

1  2  .  5  to    25 

2 

ai 

2 

I* 

18 

ai 

18 

•31 

40  to  80 

1  2  .  5  to    25 

3 

3i 

2 

ll 

20 

ai 

24 

.41 

40  to  60 

1  6  .  5  to    25 

3 

3* 

2 

li 

24 

*! 

24 

.41 

40  to  60 

1  6  .  5  to    25 

4 

4i 

2 

li 

12 

ai 

12 

•25 

40  to  80 

10      to    20 

2 

a| 

2 

*i 

14 

2* 

12 

•25 

40  to  80 

10      to    20 

2 

ai 

2 

li 

16 

ai 

18 

38 

40  to  80 

15      to    30 

2 

ai 

2 

li 

18 

ai 

18 

-38 

40  to  80 

15       to    30 

3 

si 

2 

li 

20 

a| 

24 

•51 

40  to  60 

20         tO      30 

3 

si 

2 

li 

24 

ai 

24 

•51 

40  to  60 

20      to    30 

4 

4i 

2 

ll 

12 

2| 

12 

•31 

40'  to  80 

1  2  .  5  to    25 

2 

ti 

2 

I-^ 

14 

a| 

12 

•3i 

40  to  80 

1  2  .  5  to    25 

2 

ai 

2 

I^ 

16 

** 

18 

.46 

40  to  80 

18.5  to    37 

2 

2i 

2 

li 

18 

at 

18 

.46 

40  to  80 

18.5  to    37 

3 

3i 

2 

li 

20 

at 

24 

.62 

40  to  60 

25      to    37 

3 

3l 

2 

^ 

24 

a* 

24 

.62 

40  to  60 

25      to    37 

4 

4i 

2 

li 

14 

3 

12 

•37 

40  to  80 

15      to    30 

2 

«i 

2 

li 

16 

3 

18 

•55 

40  to  80 

22       to    44 

2 

aj 

2 

li 

18 

3 

18 

•55 

40  to  80 

22       to     44 

3 

$i 

2 

li 

20 

3 

24 

•73 

40  to  60 

29       to    48 

3 

3i 

2 

li 

24 

3 

24 

•73 

40  to  60 

29      to    48 

4 

4i 

2 

li 

16 

3i 

18 

•75 

40  to  80 

30      to    60 

2 

ai 

3 

2 

18 

3l 

18 

•75 

40  to  80 

30      to    60 

3 

3i 

3 

2 

20 

3i 

24 

i  .00 

40  to  60 

40      to    60 

3 

3* 

3 

2 

24 

3i 

24 

i  .00 

40  to  60 

40      to    60 

4 

4i 

3 

2 

.    18 

-4 

18 

.98 

40  to  80 

39      to    78 

3 

3i 

3i 

2i 

20 

4 

24 

1.30 

40  to  60 

52      to  104 

3 

3* 

3i 

2i 

24 

4 

24 

1.30 

40  to  60 

52      to  104 

4 

4i 

3i 

2i 

Pumps  of  this  type  are  made  for  heavier  pressures. 


SIMPLEX    PUMPS 


181 


The  Davidson  pump  (Fig.  166)  is  operated  by  means  of 
the  oscillation  of  the  auxiliary  valve.  The  arm  A  is  attached 
to  the  piston  rod  on  the  outside  of  the  pump.  The  turning 


182  PUMPING  MACHINERY 

of  the  shaft  B  turns  the  cam  C  which  causes  the  auxiliary 
valve  D  to  turn.  When  the  piston  gets  to  the  end  of  the 
stroke  the  pin  puts  the  valve  into  the  position  shown  so  that 
the  passage  E  is  connected  with  the  steam  and  the  passage  F 
with  the  exhaust.  This  causes  the  auxiliary  pistons  GG  to  be 
driven  to  the  left  so  that  steam  will  enter  the  right  end  of  the 
main  cylinder.  The  auxiliary  valve  is  oscillated  when  it  acts 
as  an  auxiliary  valve,  but  the  same  casting  D  acts  as  the  main 
valve  by  its  longitudinal  motion.  The  casting  D  fits  into  a 


FIG.  167. — Davidson  Pump. 

cavity  in  the  frame  of  the  auxiliary  pistons  GG,  and  although 
it  is  carried  to  the  top  of  the  cylinder  in  which  GG  moves,  it 
does  not  completely  fill  the  cylinder,  so  that  the  steam  which 
enters  at  H  can  pass  around  the  valve  D.  The  exhaust  passes 
through  the  cavity  containing  the  pin  of  D  and  the  cam  C. 

The  valve  D  is  twisted  into  the  position  shown  in  the  small 
section  whenever  it  is  preparing  to  drive  the  main  valve  to  one 
end  of  the  cylinder.  When  it  is  in  such  a  position  that  E  and 
F  are  covered,  the  openings  to  the  main  cylinder  are  open  to 
either  steam  or  exhaust.  In  this  way,  when  steam  is  turned 


SIMPLEX   PUMPS 


183 


THE    DAVIDSON    PRESSURE    PUMP,    PISTON    PATTERN 
FOR  BOILER  FEEDING,  OPERATING  HYDRAULIC  ELEVATORS,  ETC. 

For  a  pressure  of   150  pounds.      (If  specially  ordered  these  pumps 
can  be  furnished  for  250  pounds  working  pressure). 

(Composition  or  bronze  cylinders  to  order.) 

SIZES    AND    DETAILS 


i.P  Boiler, 

msed  on  30 

Ibs.  of 

Size. 
No. 

Steam 
Cylin- 
der. 

Water 
Cylin- 
der. 

Stroke 
Inches 

Gallons 
per 
Stroke. 

Water  per 
H.P.perHr. 
which 
the  Pump 

Steam 
Pipe. 

Ex- 
haust 
Pipe. 

Suc- 
tion 
Pipe. 

Dis- 
charge 
Pipe. 

will  supply 

with  ease.1 

0 

2* 

ii 

3 

.022 

20 

1 

1 

1 

| 

i 

3 

if 

4 

.041 

35 

i 

1 

I 

I 

i 

35 

2 

4 

•05 

5° 

I 

\ 

il 

i 

ij 

4 

2i 

4 

.069 

65 

i 

I 

il 

i 

2 

4i 

2\ 

6 

•  13 

I25 

i 

3 
4 

l| 

il 

2* 

5 

3 

6 

•183 

i75 

8 

i 

2 

ii 

3 

Si 

3l 

8 

.28 

275 

i 

I 

2 

ri 

3^ 

6 

4 

8 

•435 

425 

1 

I 

2$ 

2 

4 

7 

4 

10 

•54 

525 

1 

I 

2^ 

2 

42 

8 

5 

IO 

•85 

850 

i 

Jl 

3 

*i 

5 

9 

5l 

12 

I  .  12 

I  IOO 

i 

il 

3 

2i 

6 

10 

6 

12 

1-47 

1400 

1 

ii 

4 

3 

7 

12 

7 

12 

2  .OO 

2OOO 

ii 

2 

5 

4 

7^ 

14 

81 

12 

2.77 

2750 

i^ 

2 

6 

5 

8^ 

14 

81 

14 

3-23 

3200 

if 

2 

6 

5 

8i 

14 

IO 

14 

4.76 

475° 

ii 

2 

8 

7 

Q 

16 

9l 

16 

4.66 

2 

2* 

7 

6 

9* 

16 

10 

16 

5-44 



2 

2 

/ 
8 

7 

IO 

18 

ioi 

18 

6.74 

2l 

•J 

8 

7 

10* 

18 

II* 

18 

*"*   *    /  T" 

8.09 

21 

O 
•J 

o 

/ 
8 

j.  w  2 
I  I 

20 

2 

III 

20 

9  .  oo 

21 

O 

y 

Q 

8 

II* 

20 

I  -2 

20 

1  1  .  49 

2* 

-J 

y 
IO 

A  •*•  2 
12 

22 

%J 
I  T. 

22 

12  .  64 

2 

0 
"2 

IO 

o 

13 

24 

O 

14 

24 

16.00 



3! 

O 

4 

10 

y 
IO 

1  Capacities  for  boiler  feeding  based  on  a  speed  of  60  single  strokes  per  minute  with 
feed  water  at  ordinary  temperature.  With  high  temperature  feed  water  reduce  H.P. 
rating  one-third.  These  pumps  can  be  run  as  slowly  as  may  be  desired,  and  in  cases  of 
emergency,  speeded  up  to  about  twice  ordinary  capacities. 

Suction  and  discharge  openings  on  both  sides.  Hand  levers  furnished 
with  sizes  No.  o  to  No.  3^,  and  with  No.  4  when  ordered.  Water-piston 
packing  and  valves  for  hot  or  cold  water  as  ordered.  Every  machine 
thoroughly  tested  before  leaving  works. 


184 


PUMPING  MACHINERY 


on  either  steam  will  enter  E  or  F  and  drive  the  main  valve  to  a 
position  to  admit  steam  to  one  end  and  start  the  pump,  or  the 
main  valve  itself  will  be  in  a  position  to  start  the  pump.  From 
this  it  may  be  seen  that  there  is  no  dead  center  possible  with 
this  pump. 

It  is  claimed  that  the  action  of  the  cam  prevents  the  piston 


FIG.  168. — Burnham  Steam  Pump. 

from  striking  the  cylinder  head.  Moreover,  the  pump  should 
start  from  any  position  and  make  its  full  stroke. 

Fig.  167  gives  an  outside  view  of  this  pump,  showing  the 
method  of  driving  the  cam  lever. 

The  table  on  the  preceding  page  gives  sizes  used  by  the  M. 
T.  Davidson  Co.  for  one  type  of  their  pump. 

In  the  Burnham  steam  pump  (Fig.  1.68)  the  auxiliary  valve 
is  placed  on  the  side  of  the  pump  and  is  driven  by  the  cam  rod 


SIMPLEX  PUMPS 


185 


A.  When  the  pump  reaches  the  position  shown  in  the  figure, 
the  rod  B  is  moved  to  the  right  and  this  moves  the  aux- 
iliary valve  H  to  the  right. 
This  valve  is  mounted  on  the 
side  of  the  steam  chest  and 
its  motion  to  the  right  will 
admit  steam  to  the  left-hand 
end  of  the  auxiliary  piston  / 
while  it  connects  the  right- 
hand  end  to  the  exhaust. 
This  drives  the  auxiliary  pis- 
ton to  the  position  shown  in 
the  figure,  and  with  it  the 
main  valve.  The  pump  is 
then  reversed. 

The  double  ports  for  the 
main  piston  and  auxiliary 
piston  are  arranged  to  prevent 
pounding  and  give  smooth 
action .  The  small  ports  lead  - 
ing  to  the  ends  of  the  cylin- 
ders are  for  the  sole  purpose 
of  admitting  steam.  The 
exhaust  takes  place  through 
the  passages  shown  full  in  the 
section  and  when  either  the 
main  piston  or  the  auxiliary 
piston  travel  over  these  ports 
the  steam  behind  them  is 
trapped  and  forms  a  cushion. 
In  the  positions  shown  in  the 
figure  each  of  the  pistons  is 
started  on  its  return  stroke 

FIG.  169. — Burnham  Pump. 

by   the    small   quantities    of 

steam  which  are  admitted  through  the  small  passages  shown 
dotted  and  extending  to  the  ends  of  the  cylinders.  Fig.  169 
illustrates  the  exterior  appearance  of  the  pump.  It  will  be 


186 


PUMPING  MACHINERY 


noted  that  the  valve  cam  moves  the  valve  rod  at  the  end  of 
the  stroke  only,  and  moreover  this  motion  is  gradual  and  slow. 
The  table  gives  the  sizes  of  these  pumps  used  for  tank 
service. 


DETAILS  OF  BURNHAM   TANK   OR   LIGHT-SERVICE    PUMPS 


a;  t-i 

™ 

0-0  , 

As  length  of  connections 

-M  U 

<A 

s.S  £ 

1 

I 

1 

increases,  larger  size 
pipes  should  be  used. 

1 

£  Z 

|l 

a«s 

D 

sa 

J5 

CO 

CO 

IH    ^ 

VnCO 

fc'£§ 

o 
g 

>> 
0 

8 

*o 

1 

j 

-M 
»•* 

§ 

is 

1 

joj 

|I 

S  *" 

bo 
.£ 

CO 

i 

J 

co 

X 

w 

I 

1 

"c3 
O 

rt  rt  O 

o 

3M 

O 

CO 

C 

8 

10 

12 

8 

6 

4.08 

204  to  408 

24,400 

50  to  ioo 

R 

8 

II 

12 

8 

6 

4-93 

247  to  494 

29,600 

50  to  ioo 

R 

8 

12 

12 

8 

6 

5-87 

293  to  587 

35,200 

50  to  ioo 

R 

8j 

6 

10 

4 

3i 

I  .  22 

73  to  147 

8,800 

60  tO  120 

R 

n 

7 

10 

5 

4 

1.66 

IOO  tO   200 

12,000 

60  to  1  20 

R 

8f 

8 

10 

5 

4 

2.17 

130  to  261 

15,600 

60  tO  120 

R 

£$ 

9 

10 

6 

5 

2-75 

165  to  330 

19,800 

60  to  i  20 

R 

8i 

10 

10 

6 

5 

3-40 

204  to  408 

24,400 

60  tO  120 

R 

10 

8 

12 

2 

5 

4 

2.61 

130  to  261 

15,600 

60  to  1  20 

R 

10 

9 

12 

2 

6 

5 

3-30 

165  to  330 

19,800 

50  to  ioo 

R 

10 

10 

12 

2 

8 

6 

4.08 

204  to  408 

24,400 

50  to  ioo 

R 

10 

ii 

12 

2 

8 

6 

4-93 

247  to  494 

29,600 

50  to  ioo 

R 

IO 

12 

12 

2 

8 

6 

5.87 

293  to  587 

35,200 

50  to  ioo 

R 

10 

14 

16 

2 

10 

8 

10.65 

404  to  808 

48,400 

38  to  75 

R 

12 

8 

12 

2- 

5 

4 

2.61 

130  to  261 

15,600 

50  to  ioo 

R 

12 

9 

12 

2 

6 

5 

3-3C 

165  to  330 

19,800 

50  to  ioo 

R 

12 

10 

12 

1 

8 

6 

4.08 

204  to  408 

24,400 

50  to  ioo 

R 

12 

ii 

12 

* 

8 

6 

4-93 

247  to  494 

29,600 

50  to  ioo 

R 

12 

12 

12 

I 

8 

6 

5.87 

293  to  .  587 

35.200 

50  to  ioo 

R 

12 

14 

12 

I 

10 

8 

8.00 

400  to  800 

48,000 

50  to  ioo 

R 

12 

8 

16 

I 

6 

4 

3-48 

132  to  264 

15,800 

38  to  75 

R 

12 

9 

16 

• 

6 

5 

4-40 

167  to  334 

20,040 

38  to  75 

R 

12 

10 

16 

1 

8 

6 

5-44 

206  to  412 

24,700 

38  to  75 

R 

12 

ii 

16 

8 

6 

6.58 

250  to  500 

30,000 

38  to  75 

R 

12 

12 

16 

j 

8 

6 

7.82 

297  to  594 

35,6oo 

38  to  75 

R 

12 

14 

16 

l 

10 

8 

10.65 

404  to  808 

48,400 

38  to  75 

R 

14 

10 

12 

i 

8 

6 

4.08 

204  to  408 

24,400 

50  tO  TOO 

R 

14 

12 

12 

] 

8 

6 

5-87 

293  to  587 

35,200 

50  to  ioo 

R 

14 

14 

12 

10 

8 

8.00 

400  to  800 

48,000 

50  to  ioo 

R 

14 

10 

16 

8 

6 

5-44 

206  to  412 

24,700 

38  to  75 

R 

14 

II 

16 

j 

8 

6 

6.58 

250  to  500 

30,000 

38  to.  75 

R 

14 

12 

16 

': 

8 

6 

7.82 

297  to  594 

35.6oo 

38  to  75 

R 

14 

14 

16 

f 

10 

8 

10.65 

404  to  808 

48,400 

38  to  75 

R 

14 

16 

16 

f 

10 

8 

14.00 

532  to  1064 

63,800 

38  to  75 

R 

14 

16 
18 

20 

12 

10 

17.40 

522  to  1044 

62,600 

30  to  60 

R 

'4 
16 

I  O 

II 

16 

8 

6 

6.58 

660  to  1321 
250  to  494 

79,200 
29,600 

30  to  60 
38  to  75 

R 

16 

12 

16 

j 

8 

6 

7.82 

297  to  594 

35,6oo 

38  to  75 

R 

16 

14 

16 

10 

8 

10.65 

404  to  808 

48,400 

38  to  75 

R 

16 

16 

16 

10 

8 

14.00 

532  to  1064 

63,800 

38  to  75 

R 

16 

16 

20 

j 

12 

IO 

17.40 

522  to  1044 

62,600 

30  to  60 

R 

16 

18 

20 

; 

12 

10 

22.03 

660  to  1321 

79,200 

30  to  60 

R 

16 

20 

24 

j 

14 

12 

32.6- 

816  to  1632 

97,920 

25  to  50 

R 

16 

22 

24 

16 

12 

39-50 

987  to  1975 

25  to  50 

R 

18 

14 

16 

\ 

31 

IO 

8 

10.65 

405  to  810 

48,600 

38  to  75 

R 

18 

16 

20 

1 

3? 

12 

10 

17.40 

52?  to  1044 

62,600 

30  to  60 

R 

18 

18 

20 

1 

3* 

12 

10 

22.03 

660  to  1321 

79,200 

30  to  60 

R 

20 

16 

20 

L 

3* 

12 

10 

17.40 

522  to  1044 

62,600 

30  to  60 

R 

20 

18 

20 

?  \ 

3* 

12 

10 

22.03 

660  to  1321 

79,200 

30  to  60 

R 

20 

20 

24 

2  \ 

3* 

14 

12 

32.63 

816  to  1632 

97,920 

25  to  50 

R 

2O 

22 

24 

* 

31 

16 

12 

39-50 

987  to  1975 

118,500 

25  to  50 

R 

"R"  signifies  removable  bronze  lining  in  water  cylinder. 
pressed  in. 


"P"  signifies  bronze  lining 


SIMPLEX  PUMPS 


187 


The  Dean  Brothers  pump  (Fig.  170)  illustrates  another 
method  of  obtaining  these  results.  As  in  the  Burnham  pump 
the  auxiliary  valve  is  placed  on  the  side  of  the  main  steam 
chest.  This  pump  differs  from  the  others  considered  in  that  the 
motion  of  the  auxiliary  valve  is  continuous  and  not  intermittent. 
This  valve  is  driven  from  a  reverse  lever  attached  by  a  link 
to  the  piston  rod,  and  the  valve  is  moved  by  a  link  at  the  upper 
end  of  this.  When  the  main  piston  reaches  the  extreme  right 
of  its  stroke  the  auxiliary  valve  is  at  its  extreme  left,  and  the 
edge  of  the  valve  uncovers  a  small  circular  port  of  the  passage 


FIG.  170. — Dean  Brothers  Pump. 

A  leading  to  the  right  end  of  the  auxiliary  cylinder.  At  this 
time  one  of  the  two  diagonal  grooves  on  the  under  side  of  the 
valve  connects  the  other  passage  B  to  the  exhaust  port  C. 
This  then  forces  the  auxiliary  piston  to  its  extreme  left, 
and  with  it  the  main  D  slide  valve,  admitting  steam  to  the 
right-hand  end  of  the  cylinder,  reversing  the  motion.  The 
auxiliary  valve  begins  to  move  to  the  right,  cutting  off  the 
steam  and  exhaust  from  the  auxiliary  valve.  This  condition 
does  not  change  until  the  valve  almost  reaches  the  end  of 
its  stroke,  when  the  "other  diagonal  groove  connects  A  to  C, 
while  B  is  uncovered  to  the  action  of  the  live  steam.  The 
auxiliary  valve  then  goes  over  to  the  other  position  and  the 


188  PUMPING  MACHINERY 

pump  reverses.  The  link  Z),  which  moves  the  auxiliary  valve 
rod,  may  be  moved  in  a  slot  closer  to  or  farther  from  the  pivot 
of  the  reverse  lever,  making  its  stroke  shorter  or  longer;  this 
makes  the  stroke  of  the  main  pump  longer  or  shorter,  as  it 
means  more  or  less  motion  of  the  main  piston  to  move  the 
auxiliary  valve  a  sufficient  distance  to  bring  the  ports  into  the 
position  for  shifting  the  auxiliary  piston. 

The  small  spindle  projecting  from  the  top  of  the  main  valve 
casing  is  used  to  shift  the  auxiliary  piston  and  with  it  the  main 
valve  which  rests  in  a  slot  in  the  auxiliary  piston.  Such  a 
contrivance  is  necessary,  as  the  main  valve  may  be  left  in  a 


FIG.  171. — Dean  Brothers  Pump. 

position  so  that  steam  could  not  enter  either  side.  This  con- 
dition rarely  occurs,  but  in  most  of  these  pumps  such  a  contin- 
gency is  guarded  against  by  some  device  like  this  or  the  pro- 
jecting rods  of  the  Marsh  pump. 

Fig.  171  illustrates  one  of  these  pumps  designed  for  boiler 
feeding,  in  which  several  special  features  can  be  seen.  The 
steam  and  water  valves  are  easily  examined  without  the  neces- 
sity of  interfering  with  any  pipes;  the  frame  is  of  steel;  the 
cylinders  are  removable  and  may  be  replaced  without  disturbing 
any  other  parts;  the  piston  rod  has  a  cross-head  fitting  over 
the  lower  rod  frame,  making  it  impossible  to  twist  the  cross- 
head  so  as  to  jam  the  valve  gear. 


SIMPLEX   PUMPS 


1S9 


The  following  table  gives  the  sizes  used  by  these  makers 
for  one  line  of  their  pumps. 

DEAN    BROTHERS    SIMPLE    PUMPS   WITH    PACKED    PISTONS 

(For  feeding  boilers  or  pumping  against  pressure.     They  will  elevate  water  200  feet  with 

60  pounds  steam.) 

TABLE    OF   DIMENSIONS 


fed 

0) 

'O 

1 

B 

1  Capacity  per 
Minute. 

£ 

$ 

i.  M 
O   C 

SSQJ 

S9 

.^H 

o 

B 

& 

o"! 

&1 

CL 

i 

£$ 
a| 
jl 

>>        • 

?j 

%* 

C 
| 

ii 

Tito 

1 

t 

£ 
13 

E  o 

xhaust  ] 
Inches. 

PL,  ^ 

II 

fl 

u-i  (^  «T 
°^  S 

w 

w 

£ 

03 

o 

& 

O 

& 

W 

oS 

Q 

•3 

C 

3           2 

4 

0.054 

2OO 

IO 

1 

| 

if 

i 

5° 

D 

4 

tf 

6 

0.14 

I4O 

20 

i 

4 

Ji 

if 

IOO 

E 

5i 

3f 

7 

•33 

125 

42 

1 

2 

220 

F 

6 

4 

10 

•55 

125 

68 

i 

Ji 

3 

2i 

400 

G 

8           5 

12 

i  .02 

IOO 

102 

i 

ri 

4 

3 

700 

H 

9           6 

I  2 

1.47 

IOO 

147 

ij 

2 

4 

3 

IOOO 

I 

10          7 

12 

2  .00 

IOO 

200 

ii 

2 

5 

4 

1300 

K 

12                 8 

12 

2.6l 

IOO 

26l 

2 

2i 

6 

5 

1600 

L 

14           9 

16 

4.40 

80 

352 

2 

3 

7 

6 

2500 

M 

14         10 

20 

6.80 

70 

476 

2 

3 

8 

7 

35°°  ' 

1  In  an  emergency  more  than  the  above  capacity  can  be  had,  but  for  continuous  work, 
such  as  feeding  boilers,  not  more  than  half  the  strokes  given  above  are  advised.  The  valve 
motion  secures  a  smooth  action  and  admits  of  regulation,  so  as  to  deliver  a  steady  supply 
of  water,  exactly  equal  to  the  amount  evaporated. 

The  pumps  described  in  this  chapter  are  taken  to  repre- 
sent this  class,  and  although  there  are  many  other  different 
forms,  the  principles  shown  in.  the  these  will  aid  in  under- 
standing the  action  of  any  other  simplex  pump.  All  of  the 
valve  gears  aim  to  move  the  valve  positively  whatever  be  its 
speed.  From  the  spring- thrown  valve  of  Worthington  to  the 
latest  form,  this  is  the  governing  idea. 


CHAPTER  V 
DYNAMICS  OF  WATER  END 

Velocity  and  Acceleration.  To  make  the  velocity  of  dis- 
charge from  a  pump  more  definite  a  pump  with  a  fly  wheel 
and  connecting  rod  will  be  considered.  The  figure  below  (Fig. 
172)  represents  a  plunger  pump  in  which  the  rate  of  discharge 
or  suction  at  any  instant  is  proportional  to  the  velocity  cf  the 
plunger.  From  the  diagram  the  movement  from  the  end  cf 
the  stroke  is  * 

x=AB+BC=r(i-cosd)+nr(i-cosa),      .     .     (i) 


FIG.  172. — Diagram  of  Pump. 

where  0  =  the  angle  the  crank  has  moved  from  its  head  dead 

center; 
a=the  inclination  of  the  connecting  rod  with  the  center 

line; 

r  =  the  crank  radius  in  feet; 

w=the  ratio  of  length  of  connecting  rod  to  crank  radius. 
Now 


. 

sin  a  =  —  sin  0. 
nr 


Hence 


cos  a 

~\ 

rip* 

*  V=I  —  \ 

?  w2 

8  0,  approx. 

Therefore 

% 

-r[i 

—  cos  0  - 

I1 

1 

v  1 

(2) 

L 

F  '  w  sm2 

J 

*  See  page  257  for  table  of  symbols. 

190 


DYNAMICS  OF  WATER  END 

dx                          sin  0  cos  6\ 
Velocity  =-jr=  S=  r  sin  0+-   

=  r  sin  0+  — sin  26  \co. 
2n 


d2 

Acceleration  =  -r 
at* 


2*        r 
p-«=r[. 


—  H  cos  0-\ 


cos  201 

n 


t>. 


-8 


30° 


-Stroke 


30°          0- 


90°    xe. 


20° 


\ 


-Stroke 


Head  End  Q  6=30°         0=60° 


191 
(3) 


•     (4) 


Craak  End 


FIG.  173. — Velocity  and  Acceleration. 

!i     . 

These  give  diagrams  such  as  shown  in  Fig.  173,  wjien  S 
and  a  are  plotted  against  x  or  piston  position  as  asbcissaeJ 

This  approximate  solution  is  sufficiently  close  for.  general 
work,  as  may  be  seen  from  the  table  below,  which  gives  the 


192 


PUMPING  MACHINERY 


FIG.  174. — Curves  of  x,  S,  a  for  Different  Crank  Angles. 


DYNAMICS  OF  WATER  END 


193 


values  of  a  by  the  approximate  formula  and  the  exact  formula 
obtained  without  expanding  the  equation  for  cos  «.  The 
exact  equations  are 


r .        sm  e  cos  e  l 

S=cor  sm#+—  

Vtt2-sin2  /9J 


COS0  + 


n2  cos2  0 


sin2# 


(5) 


(6) 


For  a  discussion  of  the  use  of  the  exact  and  approximate 
values  of  acceleration  the  reader  is  referred  to  the  papers  of 
D.  S.  Jacobus,  A.  S.  M.  E.  Transactions,  Vol.  XI,  p.  492  et  seq., 
and  p.  1116  et  seq. 

In  the  exact  formula  the  angular  velocity  is  considered 
as  constant,  and  since  this  is  not  the  case  there  is  really  an  error 
in  the  exact  formula. 

The  curves,  Fig.  174,  give  the  values  of  x,  5,  and  a  as  func- 
tions of  crank  movement.  These  curves  are  symmetrical,  and 
in  computing  the  values  of  the  points,  as  in  the  table  below, 
for  a  unit  value  of  a>  and  r,  this  symmetry  is  seen. 


• 

cos  0 

sin  d 

cos  26 

sin  20 

X 

S 

a 

a  exact 

0 

I 

0 

I 

0 

0 

0 

I  .  167 

I  .  167 

I5 

•96593 

.25882 

.86603 

•5 

.040 

.300 

I  .  I  10 

I  .  ill 

3° 

.86603 

•5 

•5 

.86603 

•J55 

•  572 

•949 

•95° 

45 

.70711 

.70711 

o 

i 

•335 

.790 

.707 

.706 

60 

•5 

.86603 

-  -5 

.86603 

•563 

•  938 

.417 

.417 

75 

.25882 

•96593 

-  .86603 

•5 

.819 

1  .007 

.114 

•**J 

90 

0 

I 

—  i 

o 

1.083 

1  .000 

-.167 

—  .  169 

i°5 

-  .25882 

•96593 

-  .86603 

-  -5 

i-346 

.924 

-  .403 

-  -405 

I2O 

-  -5 

.86603 

-5 

—  .86603 

1.562 

•794 

-.583 

-.583 

135 

-  .70711 

.70711 

o 

—  i 

1-749 

.624 

-.707 

-  .709 

J5° 

—  .86603 

•5 

•5 

—  .86603 

i  .908 

.428 

-•783 

-.782 

165 

-  -96593 

.25882 

.86603 

-  -5 

1.972 

.217 

-  .822 

-  .821 

180 

—  i 

0 

i 

0 

2 

o 

-•833 

-•833 

To  get  actual  S  the  values  are  multiplied  by  ra)lt  a  by  ra>2,  and  x  by  r. 

Discharge  and  Discharge  Curves.  The  momentary  rate  of 
discharge  from  any  pump  will  be  A  X  velocity,  where  A  rep- 
resents the  effective  area  of  the  piston  or  plunger.  A  has 


194  PUMPING  MACHINERY 

various  values  for  the  different  types  of  pumps.      If  D  =  diam- 
eter of  cylinder  and  d  =  diameter  of  rod,  the  following  results: 

TtD2 

CASE  i.  For  the  plunger  pump,  A  = . 

4 

For  the  piston  pump,  on  end  without  rod, . 

CASE  2.  ,      4    , 

For  the  piston  pump,  on  end  with  rod,  —  ( D2  —  d2 } . 

4  \  / 


CASE  3. 


TO 

For  the  bucket  pump  down  stroke,  — . 

4 

For  the  bucket  pump  up  stroke,  -f/)2—  d2}. 


Differential  pump,  down  stroke,  -— . 

r»i  '"  ••  .';  4 

Differential  pump,  up  stroke,-  -( D2  —  d2  j .      (See  Fig.  228.) 
The  discharge  then  becomes 

^';     K  Total  Q-fdQ-£A%dt,     -:-_.V.    •     (7) 

Q  =A  I   dx=A  X  distance  passed  over, 

=  ALN'.     .    .    .     .     .     ,     .     ,     .     .     .  '  .     (8) 

L  =  length  of  stroke, 
N'= effective  pumping  strokes  in  given  time. 

The  action  of  the  discharge  or  suction  may  be  shown  by  a 
diagram,  Fig.  175,  in  which  abscissae  represent  time  and  ordi- 

nates,  the  quantity  A-rr.    The  area  of  this  curve  is  Q  and  the 
mean  height  of  the  curve  would  be 

**$Q-4!f-ALS» (9) 

N"  =  effective  strokes  per  second. 
To  show  the  variation  of  the  momentary  values  from  the 


DYNAMICS  OF  WATER  END 


195 


30    GO    90     120    150   180    210    240    270    300    330 


30    60    90    120    150   180    210    240   270    300    330   360 


0     30    60     90     120    150   180     210    240   270    300    330   360 

CaseS 


120    150    180    210    240    270    300   330    360 


0     30    60    90    120    150   180    210    240    270    300    330   860 

Case  5 

FIG.  175. — Pump  Discharges. 

mean,  the  mean  value  may  be  plotted  (Fig.  175 )  as  a  dot  and 
dash  line  on   the   diagram.     Two  additional  cases  have  been 


196  PUMPING  MACHINERY 

added:  One,  Case  4,  in  which  there  are  two  double-acting  pumps 
and  another,  Case  5,  where  three  single-acting  pumps  are  used. 
The  series  of  diagrams  are  drawn  for  pumps  in  which 
the  diameter  of  the  pump  barrel  is  20  inches;  the  diameter 
of  the  piston  rod,  where  used,  2  inches;  the  diameter  of  the 
equalizing  plunger  rod,  where  used,  14  inches,  and  the  stroke 
30  inches.  The  speed  of  the  pumps  will  be  taken  at  sixty 
revolutions  per  minute  and  n=6.  The  quantities  discharged 
per  second  for  the  various  cases  are  as  follows: 


.16 

CASE  I.    q=  —  xLxN"=-~  X-Xi  =5-4475- 
4 


CASE  2.    q=\—  +-(D*-d*)\xLxN 
L   4       4  J 

[314.16    311. 
+ 


CASE  3.    2=  — 
4 


xd2 

— 

4 


144 

*)  =1^  =2.160. 
144 

.16 


CASE  4.    q  =  zq  of  Case  2.  » 

Ah  and  Ac  same  as  Case  2. 
CASE  5.     q=$q  of  Case  i. 

A  =same  as  Case  i. 

The  curves  show  clearly  the  way  the  quantity  varies  as 
the  pump  moves.     The  variation  due  to  the  piston  rod  is 


DYNAMICS  OF  WATER  END 


197 


shown  in  the  figure  by  the  different  heights  of  the  curve  on 
the  forward  and  back  stroke  of  the  double-acting  pump.  The 
effect  of  a  number  of  cylinders  on  the  uniformity  of  flow  in 
the  discharge  pipe  is  seen  by  comparing  the  curve  from  the 


Li  and  L-2    represent  length 
of  pipes; 

h2  and  h'2  represent  vertical 
heights  of  lift. 


FIG.  176. — Pump  Arrangement. 

single-cylinder  pump  with  the  resultant  curve  from  the  multi- 
cylinder  pump. 

Forces  on  Piston.      The  velocities  in  the  discharge  pipe  arid 
suction  pipe   are   dependent   on  the  velocity  of  the  piston,  as 


198  PUMPING    MACHINERY 

water  may  be  considered  as  incompressible.  Since  this  will 
mean  a  variation  in  pressure  due  to  the  change  of  momentum, 
it  is  necessary  to  consider  the  forces  acting  in  the  suction  pipe 
and  delivery  pipe. 

Consider  one  side  of  the  pump,  Fig.  176,  having  a  piston 
area  A  and  a  stroke  L.  To  make  this  a  general  case  the  cylinder 
will  be  inclined  at  an  angle  <£  to  the  horizontal.  The  dimen- 
sions L  represents  lengths,  while  A\,  A  2  are  the  areas  of  section 
of  the  various  channels  through  which  the  water  passes,  and  h, 
the  pressures  measured  in  feet  of  water  at  the  various  points. 

The  first  pressure  necessary  to  find  is  that  acting  on  the 
lower  side  of  the  piston  during  an  upward  stroke.  This  may 
be  written  as 

Sm  0),       .       .       ..     (lo) 


h8=  pressure  on  suction  side  of  piston  expressed  in  feet; 
/*a=head  corresponding  to  the  pressure  of  the  atmosphere; 
h2=  height  to  end  of  piston  stroke  from  water  level; 
A3=  friction  and  other  losses  except  that  of  inertia; 
A4=  force  due  to  inertia  expressed  in  feet  head; 
/rs.v.  =  friction  in  suction  valve; 

z'sin  <£  =  the  vertical  movement  of  the  piston. 

The  pressure  hs  is  the  mean  pressure  on  the  piston  and  the 
actual  pressure  at  the  center  of  the  piston;  there  is  a  slight 
variation  over  the  area. 

Losses  in  Pipes.  The  friction  losses  in  the  suction  side 
are  made  up  of  three  principal  parts:  That  in  the  suction 
pipe,  that  in  the  valve  boxes,  and  that  in  the  passages  of 
the  pump.  The  general  formula  for  -this  loss  at  any  instant, 
as  given  in  works  on  theoretical  hydraulics,  is 


or  using  later  experimental  work, 

h  =  klvn.  ....      (12) 


DYNAMICS    OF    WATER    END 


199 


The   velocity  of   the  water   in   any  member   of   the  pump 
varies  in   the   same  manner  as  that  of  the  pistcn  provided  the 
water  stream  does  not  separate  but  follows  the  piston.     Hence 
a  mean  value  of  the  loss  per  cubic  foot  must  be  obtained. 
Since  v2  varies,  the  mean  loss  per  cu.ft.  is 


2g 


i 

--  dq 
o    di  2g  a 


q  q 

since  dq=AS  dt,  where  5  =  the  speed  of  the  piston. 
Now  v\  = — .     This  gives 
Al, 


To  integrate  this  for  any  given  pipe  the  values  of  the  coeffi- 


FIG.  177. — Curve  of  fS3. 

cient  /  for  different  velocities  v\  in  the  pipe  are  found,  and  the 
product  fS3  is  found  at  different  intervals  of  time,  or  what  is 
the  same  thing,  for  different  values  of  6  and  are  plotted  on  a  0 
base.  This  gives  a  diagram  shown  in  Fig.  177. 

The  area  of  this  figure,  found  by  a  planimeter  or  calculated 
by  Simpson's  Rules,  is  the  value  of  the  integral  and  from  this 
the  mean  h  loss  may  be  found. 


200  PUMPING    MACHINERY 

The  instantaneous  value  of  the  loss  at  any  point  per  pound 
reduces  to 


This  form  of  an  expression  holds  for  the  loss  in  the  pas- 
sages leading  from  the  valve  to  the  cylinder  and  through  the 
valve,  although  there  is  also  a  direct  loss  due  to  the  pressure 
required  to  hold  up  the  suction  valves.  The  form  will  also 
represent  the  losses  at  entrance  to  the  pipe,  and  other  obstruc- 
tion. 

The  values  for  these  various  terms  will  now  be  discussed. 

For  many  years  it  has  been  customary  to  consider  that  the 
loss  of  head  in  a  pipe  line  was  given  by  Eq.  (n), 


where  hp=\oss  of  pressure  expressed  in  feet; 

/=a  coefficient  which  varies  with  the  kind  of  pipe,  diam- 

eter, and  velocity; 
/=  length  of  pipe  in  feet; 
v=  velocity  in  feet  per  second; 
d=  diameter  of  pipe  in  feet; 
g  =  acceleration  of  gravity  in  feet  per  sec.  per  sec. 

Since  /  varies  with  the  velocity  an  attempt  was  made  to 
eliminate  from  this  term  the  velocity,  and  for  this  purpose  the 
equation  was  put  in  the  form  Eq.  (12): 

h=kvnl  ......     .      (12') 

By  plotting  the  results  of  many  experimenters  on  log- 
arithmic paper  and  by  the  methods  of  least  squares  Reynolds 
and  others  have  found  that  n  has  a  value  of  i  for  low  velocities. 
The  value  of  n  changes  to  about  1.75  when  the  velocity  exceeds 
a  certain  amount.  The  point  at  which  this  change  occurs  is 
called  the  critical  velocity.  Reynolds  showed,  by  injecting  a 
colored  liquid  in  a  fine  stream  into  the  water  in  a  glass  tube, 
that  so  long  as  the  velocity  was  less  than  the  critical  velocity 


DYNAMICS  OF  WATER  END  201 

the  colored  stream  remained  as  a  thread,  showing  that  the 
v  ater  was  traveling  in  stream  lines;  but  as  the  critical  velocity 
vras  reached  the  color  became  mixed  with  the  other  water, 
showing  that  there  were  eddies  throughout  the  whole  stream. 
Some  have  found  this  same  effort  by  other  means. 

If  the  results  of  a  great  number  of  experimenters  are  exam- 
ined, n  will  be  found  to  have  various  values,  varying  from  1.7 
to  2,  rough  pipe  seeming  to  give  the  higher  value.  The  value 
of  k  of  the  formula  depends  on  d,  and  F.  C.  Lea  in  his  "  Hy- 
draulics "  shows  that 

k=kd-ij*. 

This  gives,  then, 

,    k'iv1-75          k'lv™ 


to 


There  are  thus  two  formulae  giving  the  loss  in  a  pipe  line, 
one  in  which  the  loss  is  expressed  in  terms  of  the  square  of  the 
velocity  and  a  factor  which  varies  with  the  diameter  and  velocity, 
and  another  in  which  the  loss  is  expressed  in  terms  of  powers 
of  the  velocity  and  diameter  and  a  factor  which  is  a  constant 
for  one  kind  of  pipe. 

For  general  purposes  in  cast-iron  pipes  the  loss  given  by  the 
two  formulae  is  by  the  approximate  forms  : 


For  ordinary  cast-     "-^UT  —         •     •     •     (J5') 


iron    pipe,    not 
clean 


V2 

hs  =  0.00055^ 


The  value  of  /  for  clean  pipe  is  given  in  tables  in  books  such 
as  Merriman's  "  Hydraulics  "  and  Lea's  "  Hydraulics."  The 
tabular  value  for  /  is  about  0.02  for  clean  pipe  when  an  average 
value  only  is  needed.  For  an  exact  value  when  the  table  is 
net  at  hand  the  following  may  be  used  for  clean  pipe: 

,    0.026 


202  PUMPING  MACHINERY 

-,2 

The  two  expressions  for  Joss  may  be  put  in  the  form  hs  =  £  — 


where  £  has  the  value/-  or    125  25.     In  each  case  £  has  a  value 


7 
-         125 

which  varies  with  the  velocity  to  a  small  degree,  and  as  the  other 
losses  are  expressed  in  terms  of  fl2,  it  is  well  to  express  this  loss 
in  the  same  manner. 

The  less  at  entrance  into  a  pipe  line  depends  on  the  arrange- 
ment of  the  end.  If  a  plain  end  flush  with  the  wall  of  the  fore 
bay  is  used  the  loss  at  entrance  has  been  shown  (Merriman's 

V2 

"  Hydraulics/'  page  207)  to  be  0.49-^,  while  if  the  pipe  pro- 

2£ 

jects  through  the  wall  the  coefficient  becomes  0.93,  and  if  a 
mouthpiece  is  used  on  the  end  of  the  pipe  the  coefficient  is 
very  small.  In  general  it  may  be  taken  that  the  loss  at  en- 
trance is  given  by  the  equation 


*.=- 


The  loss  due  to  bends  and  obstructions,  such  as  valves,  is  given 
in  the  form 

7          z_r  I"  /       \ 

h  =  kf—,     .......     (19) 

d2g 

for  bends,  and 

h  =  m  —  ,  .     .    .   -v    .    v    .    »     (20) 

2£ 

for  valves  and  cocks,  where  k  is  a  multiplier  of  the  /  of  straight 
pipe  and  has  values  depending  on  the  ratio  of  the  radius  of 
the  bend  to  the  radius  of  the  pipe  while  /  is  the  length  of  the  center 
line  of  the  bend,  m  depends  on  the  amount  the  valve  or  cock 
is  closed. 

The  values  of  k  and  m  for  the  different  cases  are  given  in 
the  forms  of  curves  (Figs.  178,  179,  180).  These  have  been 
constructed  from  the  reported  results  of  Williams,  Hubbel 
and  Fenkell,  Weisbach  and  Grashof.  In  these  R  =  radius  of 
pipe  bend  of  90°,  d  =  diameter  of  pipe,  the  quantity  df  =  amount 


DYNAMICS  OF  WATER  END 


203 


k  - 

2.0  4  

t 

1.8  ^-  

\, 

1.6  -  ^  

__sv_ 

V^ 

1.4  ----------  -----I~I-~  ----!--  --L. 

1.2  

0                        5                       10                      1 

5       ,  FU         20                      25                      30 

V  rl  ' 

FIG.  178. — Values  of  k  for  Loss  in  Bends.  . 


10 
100 


FIG.  179. — Values  of  m  for  Gate  Valves. 
(For  Upper  Curve  use  Lower  Values.) 


12 
120 


14, 
140 


204 


PUMPING  MACHINERY 


80  120  160  300  240 

FIG.  180. — Values  of  m  for  Cocks  and  Butterfly  Valves. 

Upper  Curve — Cocks. 

Lower  Curve — Butterfly  Valves. 

the  valve  has  been  closed,  and  6  the  angle  the  cock  or  butter- 
fly valve  has  been  turned.  Fig.  181  shows  the  form  of  valves 
used. 

The  loss  due  to  sudden  enlargement  which  occurs  when  the 


FIG.  181.  —  Gate  Valve,  Cock  and  Butterfly  Valve. 

water  enters  the  cylinder  from  the  valves  or  from  the  suction 
pipe  is  equal  to 

h     fa-")2 


in  which  vs  is  the  velocity  before  enlargement  and  v  the  velocity 
after  this  occurs. 


DYNAMICS  OF  WATER  END  205 

The  loss  due  to  sudden  contraction  is  much  less  for  the 
same  change  of  section,  and  is  equal  to 

2,,  2 


where 


,  0-          .  .  ,  x 

£=0.582+-—^-,    ......   (23) 


^  =  ratio  of  the  diameter  of  the  small  pipe  to  that  of  the  large  pipe. 

The  losses  of  pressure  include  those  of  velocity  heads,  and 

since  the  water  has  finally  a  velocity  5  in  the  pump,  the  velocity 

S2 
head,  —  ,  must  be  one  of  the  terms  of  the  reduction  in  pressure, 

2g 

as  there  was  originally  no  velocity  in  the  fore  bay  or  place 

AS 
from   which   the   water  is   drawn.     The   initial   velocity,  —  —  , 

AI 

in  the  suction  pipe  is  equivalent  to,  and    requires  a  pressure 

i  S2 

—  ,    which   is   greater    than   —  .     This   greater  velocity 

2£  2g 

head,  however,  is  changed  into  pressure  head  as  the  velocity 
decreases,  so  that  when  the  piston  is  reached  the  only  part 
of  velocity  head  left  which  requires  a  pressure  from  the  piston 

.    S2 
is  —  . 

2g 

Lach  less  may  be  reduced  by  means  of  the  equation 


so  that  each  term  is  expressed  as  a  function  of  S2. 
Each  equation  reduce's  then  to  the  form 

A 


These  various  equations  give  the  values  £  for  each  one  of 
the  losses,  and  hence  the  expression 


206  PUMPING  MACHINERY 

may  be  found.     This  represents  the  variation  in  head  due  to 
the  pipe  line. 

Water  Inertia  Force.  The  loss  h±  is  computed  by  considering 
the  inertia  of  the  water. 

The  masses  of  water  in  the  various  parts  of  the  pump  are 
as  follows: 

In  the  pump  Aocw  in  pounds. 

In  the  pump  spaces  A2L2w  in  pounds. 

In  suction  pipe  A\L\w  in  pounds. 

It  is  to  be  remembered  that  L  designates  the  length  of  the 
pipe  and  not  the  lift  necessarily. 

These  masses  when  multiplied  by  the  acceleration  will  give 
the  force  required  to  accelerate  them  in  poundals,  while  by 
dividing  by  g  the  result  is  in  engineering  units  of  pounds.  It 
is  to  be  noted  that  the  acceleration  of  the  piston  is  felt  in  all 
of  the  water  in  the  system,  and  is  inversely  proportional  to 
the  areas,  as  in  the  case  with  the  velocities.  From  these  the 
following  formulae  may  be  derived,  remembering  that  the  force 
is  exerted  over  the  areas,  A,  AI,  A2,  etc.: 

Axwa  ,    , 

-  ;       .     .     .     ...    ....     (25) 

6 

A2L2wa2     A2L2wAa  ,   ., 

—  =^--j  -  ;      ....     (26) 


g  gA2 

A\L\wa\     A\L\wAa 


,    N 

-    •  •  •  ••  .(27) 


These  are  typical  terms,  and  if  there  is  a  series  of  changes 
in  the  lines  these  reduce  to 


(28) 

g\  ^2  -AI/ 

where 

x  =  r\  i  —cos  0+—  sin2  0   ; 
L  2n  J 

5  =  roj  sin  0+—  sin  20   ; 
L  2n  J 

a  =  ru>2\  cos  0H —  cos  20  . 

» 


DYNAMICS  OF  WATER  END  207 


This  total  term,  h\,  may  be  divided  into  two  parts,  one  - 


and  the  other    z        + 

\    Ai          A2/g 

The  value  of  xa  is  found  by  multiplying  the  expressions  for 
x  and  a  in  Eqs.  (2)  and  (4)  or  by  multiplying  the  values  of  x 
and  a  if  these  have  been  computed  as  on  page  193,  where  the 
value  of  n  is  6. 

By  the  first  method: 


xa 


=  im  i  -cos  0+—  sin2  6     cos  0+-  cos  26 
I  2n          J  I.  n  J 

=  ^2r2]  cos  6  —  cos2  0H —  sin2  6  cos  0H —  cos  20 —  cos  0  cos  20 
L  2^  n  n 

+-i- sin2  0  cos  20  . 

2W2  J 

The  third  and  last  terms  of  this  could  be  omitted,  but  the 
other  terms  are  of  importance;  thus  the  approximate  value 
would  be 

xa  =  &2r2[cos  0(i— cos  0)  +  -  cos  20(i— cos  0)] 

n 

=  0)V[cOS0-f-  COS20][l-COS0].      ......       (29) 

By  the  second  method  for  n  =  6,  the  values  are: 


e= 

0            15 

30 

45 

60 

75 

90 

xa 

tfV2~ 

0.044 

0.147 

0.237 

0-235 

0.093 

-0.181 

6= 

105 

1  2O 

135 

150 

165 

180 

xa 

-0.592 

—  O.9II 

-1.236 

-1-494 

—  1.621 

-1.667 

W2r2 

These  values  can  be  used  to  find  the  inertia  force  at  any 
position  of  crank  and  piston  for  a  given  pump,  determining 
the  quantity  h±  of  equations  (10)  and  (28). 

Valve  Losses.  One  term  to  require  explanation  is  the 
resistance  hsv  of  the  suction  valve.  The  discussion  following 
is  due  to  Hartmann  and  Knoke  and  is  based  on  the  work  of 
Bach. 


208 


PUMPING    MACHINERY 


There  are  three  types  of  valves:   lift  or  beat  valves,  clack 
valves,  and  slide  valves,  as  shown  in  Figs.  182,  183,  and  184. 


FIG.  182.—  Lift  Valve.        FIG.  183.—  Clack  Valve.       FIG.  184.—  Slide  Valve. 

The  valves  serve  the  purpose  of  alternately  connecting  the 
pump  cylinder  to  the  suction  or  discharge  pipe  and  with  the 
exception  of  the  last  form  of  valve  they  are  operated  by 
the  water  of  the  pump.  The  slide  valve  is  operated  by  some 
external  means  as  an  eccentric.  At  times,  however,  the  clack 
or  lift  valve  may  be  operated  in  part  by  some  positive  gear. 

Now  the  resistance  of  the  passage  of  water  through  valves 
as  investigated  by  Bach  is  divided  into  two  parts:  (a)  that 
due  to  the  opening  of  the  valve,  and  (b)  that  due  to  the  passage 
of  the  water  through  the  valves. 

Pressure  Drop  at  Opening  of  Valves.  The  valve  has  the 
velocity  and  acceleration  of  the  water  under  it,  the  latter  being 

A 

a-;-,  where  Ap  is  the   area   of   the  passage  beneath  the  valve. 
Ap 

W  a  A 
If   the  valve  weight   is   Wv  its   inertia  will  be  -  -  -j-  .     If  the 

6    ^-P 

total  spring  pressure  is  taken  as  Fs  and  the  area  of  the  upper 
side  of  the  valve  is  Av,  the  following  equation  exists  between 
pressures  above  and  below  the  valve,  when  valve  is  just  opening: 


6 


pi—pu  =  loss  in  pressure  in  opening  valve. 
A,-A\  .W.F.     W,aA 


DYNAMICS    OF    WATER    END  209 

On  the  upper  side  of  the  discharge  valve  the  following 
relation  holds: 


where  ha  represents    the  atmospheric  pressure,  Hd  the  static 

pressure  in  the  pipe  line,   and  -  -  -j-a  the  force   required  to 

£  **d 

accelerate  the  water  in  the  discharge  pipe. 

This  value  may  be  determined  in  any  given  case   before 
finding  the  quantity  hdv  given  below  for  loss  in  delivery  valve. 

This  may  be  expressed  in  head  as 


For  the  suction  valve  this  equation  becomes: 

TT/"  A 

pi  A  p  =  puA  v  +  Wv  +  F8  +  ~<TT-. 

6       ^-P 

Wv    F.      Wv    A 
A-+Tp+-iaA^\'      •     (33) 

an  expression  exactly  similar  in  form.     In  this^case,  however, 
it  is  better  to  find  pi  originally  by  the  formula: 

.      y  <.      .      (34) 

to  suction  valve,  i.e.,  lift,  losses  and  velocity 
head,  and  L  =  length  to  this  point)  and  then  pu  by  the 
formula: 

W   A 

Appl~Avpu  +  W.+F,  +  -^-ra.     .    .    .    (35) 

6      ^P 

Then 

++         -  w 


Before  numerical  values  can  be  found  for  these  expressions 
the  value  of  Fs  must  be  determined.  The  Eqs.  (32)  and  (36) 
show  the  dependence  of  the  pressure  required  to  open  the 


210  PUMPING   MACHINERY 

valve  on  the  pressure  above  the  valve,  the  weight  of  the  valve, 
the  spring  pressure,  and  the  acceleration  which  depends  on  the 
speed  of  the  pump  in  revolutions  per  minute.  The  resistance 
or  pressure  increases  with  all  of  these.  The  effect  of  area  in 

-^^3 •  KvW  AkAui       I!       ...ju^  A<A___xl>.V%- .      '  ~3w» A"l  "Ivvv 

MSSSSl         {^^^^^m_^         ^^^^^x^^^ imm> 

i  ii  in  iv  y 

FIG.  185. — Valve  Forms. 

decreasing  the  value  of  this  resistance  shows  what  an  impor- 
tant part  the  valve  area  plays  in  the  design  of  pumps.  The 
pressure  drops  considered  above  occur  as  valve  is  just  opening. 
Valve  Friction.  The  resistance  of  valves  to  the  flow  of 
water  beneath  them  during  the  open  period  of  the  valve  has 
been  found  to  be  given  by  the  expression 


(37) 


where  v\=  velocity  of  water  in  passage  leading  to  valves;  hv  = 
the  loss  in  feet  head  while  valve  is  open;  ^  =  a  coefficient,  the 
value  of  which  is  given  by  formulae  below,  depends  en  the 
form  of  valve  and  has  been  investigated  for  the  forms  shown 
in  Fig.  185  in  which  Form  I  is  a  plain  disc  valve  with  no 
guides,  Form  II  is  a  disc  with  guide  vanes  attached  to  the 
lower  face,  Form  III  the  conical  valve  with  no  lower  guides, 
Form  IV  a  complete  conical  valve,  and  Form  V  a  spherical  valve. 


,  for  valves  of  Forms  I  and  IV.      .    .     (38) 
for  valves  of  Form  IL      •     •     (39) 


\  2 

;  =  a  +ft-£  +  r(-±)  ,  for  valves  of  Forms  III  and  V.     (40) 

Bach  gives  the  various  values  for  the  constants  and  pro- 
portions for  the  parts  which  reduce  to  the  forms  below: 


DYNAMICS  OF  WATER  END  211 

For  Form  I: 

i,    dl  +    dl 
h  =  —  to  —  ; 

10       4 


«  =0.55  +4 


/?  =  o.i5  to  0.16. 

h  =  rise  of  valve  in  feet; 
di  =  diameter  of  passage  leading  to  valve; 

bi  =  breadth  of  seat  in  feet  =—  to  —  •  ; 

10        4' 

i=  number  of  ribs; 

s  =  thickness  of  ribs  at  end. 

For  Form  II: 

d,       d, 
h  =  ^  to  -; 
«        4 


10       4 

a  =  a  for  Form  I  multiplied  by  0.8  to  1.6; 
/?  =  i.7o  to  1.75. 


For  Form  III: 

di       di 

h=  —  to  —  ,  &i  = 
10        4 

a  =2.6; 
/?=-o.3; 
^=0.14. 
For  Form  IV: 

L    dl  fn  ^1. 
k=~8  134' 

a  =0.6; 
/?  =  o.i5; 
For  Form  V: 

dl       dl 
h=  —  to  —  ; 
TO        4 

a  =2.7; 
/?=-o.8; 


212  PUMPING  MACHINERY 

The  values  of  these  coefficients  were  determined  when  the 
clear  space  between  the  valves  was  1.8,  the  area  of  the  passages 
between  the  valves,  or 


d2  =  diameter  of  space  to  point  midway  between  valves, 
d  =  diameter  of  valve,  and 
di  =  diameter  of  passage. 

With  these  coefficients  the  loss  through  the  valve  may  be 
determined  if  the  value  vi  is  known.  This  will  vary  with  the 
speed  of  the  plunger. 

The  discharge  area  beneath  a  valve  changes  as  the  valve 
lifts,  and  it  may  change  so  that  the  velocity  under  the  valve, 
designated  by  c,  may  remain  constant. 


(41) 


where  A  =area  of  piston  or  plunger; 

5=  speed  of  plunger  at  any  instant; 
^4p=area  of  passage  beneath  valve; 
vi  =  velocity  of  water  in  passage  beneath  valve; 
T  =  coefficient  of  discharge; 
v  =  velocity  of  discharge  radially; 
d=  outside  diameter  of  valve; 
h  =  lift  of  valve. 

In  this   case,  however,  the   movement   of   the   valve  itseif 
will  cause  a  change  in  the  velocity  v  or  in  the  height  h,  because 

r.d2 
a   quantity  of  water  —  02  ,  where  V2  is  the  velocity  of  the  valve, 

is   held  beneath   the   valve   during   the   lifting   of  it  or  is  dis- 
charged on  the  dropping  of  the  valve.     This  then  gives 


2,     ....      (42) 
calling  xd  =  l, 

.AS  =  APvi  =  rvlh±AvV2  .....      (43) 

(+  during  rise  of  valve;    —  during  fall.) 


DYNAMICS   OF    WATER   END 


213 


This   means   that   the   motion   of  the  water  in  the  passage 
below  the  valve  relative  to  the  valve  is  vi±V2,  and  therefore 


b  _ 
c~ 


/     N 
(44) 


Hartmann  and  Knoke  investigated  this  loss  during  the 
stroke  of  a  pump,  finding  the  values 
of  h  at  different  parts  of  a  stroke 
experimentally  by  an  indicator  whose 
pencil  motion  was  attached  to  the 
valve.  From  this  they  computed  £ 
by  Eq.  (38)  at  various  points.  They 
then  computed  the  velocity  of  the 
water  c  and  from  the  curve  of  move- 
ment of  the  valve  determined  v%.  The 
computation  of  h  showed  that  the 
head  over  a  large  part  of  the  stroke 

FIG.  186. — Valve  Friction. 

was  practically  constant,  the  variations. 

being  at  the  end.  The  curve  found  was  similar  to  Fig.  186. 
This  showed  that  the  loss  could  be  computed  as  if  the  water 
were  moving  at  the  speed  determined  by  the  crank  position 
of  90°  with  a  zero  velocity  of  the  valve  or  at  a  point  where  the 
valve  has  reached  the  top  of  its  stroke. 
Here 

yioQo      Atoff  .     ^ 

-J-. (45) 


A 


and 


(46) 


Size  of  Valves.  To  determine  the  vajues  of  £  and  hv  and 
also  to  find  hsv  and  hdv  the  size  of  the  valve  or  the  number  of 
valves  of  a  certain  size  for  a  given  pump  must  be  known.  The 
following  discussion  may  be  used  to  determine  the  size  or  num- 
ber of  the  valves  used  on  a  pump. 

Now  near  the  center  of  the  stroke 


Apw 


(47) 


214  PUMPING  MACHINERY 

where  Fs  is  the  spring  tension,  Wv  is  the  weight  of  the  valve, 
and  hc.  is  the  resistance  from  the  valve  expressed  in  feet  of 
water.  From  this 


(48) 

^^    C    ^  ^~Lp(JU 

Now, 


where  f  is  the  coefficient  of  discharge  beneath  the  valve  and  v 
is  the  radial  velocity  through  the  valve  opening  of  lift  h. 


flh 


or 


i          W,+F,A9  .    . 

7=\/2£  —  A~     ~77  >      •    •   •    (49) 


or  using  the  area  Av  of  the  valve  and  a  new  coefficient  by 
analogy  this  becomes 


It  is  to  be  noted  that  this  is  expressed  in  terms  of  feet  head 
of  water  exerted  over  the  complete  valve  area  Ac. 
The  equation  then  given  for  h  is 

h  = 


T^    /    W,+F.' 

V^\2g   Aw 
or 

B.^ 

.........         (51) 


where  b  expresses  the  feet  of  water  equivalent  to  the  pressure 

Y 

on  the  whole  area  of  the  valve  and  n  =  ~y=-. 

V  tt 

Since 


or 

n>=/A/2£&  .......   .   (52) 


DYNAMICS  OF  WATER  END  215 

This  formula  is  true  for  a  definite  velocity  v  and  a  definite 
spring  force  F  for  the  given  compression.  It  is  necessary  to 
investigate  the  value  of  /x  from  actual  tests.  Equation  (43) 
becomes,  at  the  end  of  the  stroke, 


fv         A       ird2 


I  ho     4irdho     4/^0* 

In  this  ho  is  the  height  of  the  valve  from  the  seat  at  the  end 
of  the  stroke.  Showing  that  at  the  end  of  the  stroke  the  valve 
is  not'  on  its  seat.  This  might  permit  slip  to  occur  From  (52) 


is  value  of  b  at  this  time 

4-  ......    ...    (S3) 

V2  4«0 

If  V2  is  assumed  to  vary  with  bo  it  is  evident  that  /*  will  vary 

4-. 

4^0 

Hartmann  and  Knoke  measured  the  movement  of  the  valves 
of  a  certain  pump  with  given  springs  by  means  of  an  indicator 
and  in  this  they  could  measure  the  quantities  of  Eq.  (43)  to 
find  pi.  By  measurements  they  obtained  b  and  using  this  and 
the  value  of  "p  (52)  the  value  of  M  could  be  found  and  plotted 

against  —  of  different  speeds.     The  results  of  this   are  seen  in 
4« 

Table  I  below  and  Fig.  187. 

It  is  seen  that  /*  has  a  maximum  value  at  about  -7=25. 

4# 

There  is  little  variation  from  this  when  -7  =  50.     This   latter 

value  gives  a  va'ue  of  M  slightly  different  from  the  previous 
value,  but  does  give  a  value  of  h  one-half  as  large  as  before  for  a 
given  value  of  d.  This  is  of  importance  in  fixing  the  sizes  of 
values  for  initial  discharge.  In  general  ho  is  made  -^d  and 
for  this  n  =  0.845  or  °'85« 


216 


PUMPING  MACHINERY 


TABLE   I 

VALUES   OF 


h 

d 

k 

d 

mm. 

ins. 

4* 

ti 

mm. 

ins. 

.  4h 

in 

o.o 

0.000 

o  .  650 

6.0 

0.236 

2.50 

°-532 

.  i 

.004 

150  .OO 

.710 

6-5 

.256 

2.31 

•523 

.  2 

.008 

75.00 

.780 

7.0 

.276 

2.14 

•3*3 

•3 

.012 

5°  •  °° 

•845 

7-5 

•295 

.00 

•5°7 

•4 

.016 

37-5° 

.890 

8.0 

•315 

.87 

•  5°° 

•5 

.020 

30.00 

.911 

8-5 

•335 

.76 

•493 

.6 

.024 

25.00 

•913 

9.0 

•354 

.67 

•485 

.8 

.031 

i8-75 

.902 

9-5 

•374 

•58 

•477 

I  .0 

•039 

15  .00 

.870 

IO.O 

•394 

•5° 

.472 

J-5 

•059 

10.00 

.788 

ii  .0 

•433 

•36 

•459 

2  .O 

.079 

7-50 

•732 

12  .O 

'•472 

•25 

•445 

2-5 

.098 

6.00 

.690 

I3.0 

•512 

•i5 

•431 

3-o 

.118 

5  -°° 

.650 

14  .  o 

•551 

.07 

.420 

3-5 

•138 

4.28 

.622 

I5-° 

•591 

.00 

.407 

4.0 

•J57 

3-75 

•599 

16.0 

.630 

•94 

•395 

4-5 

.177 

3-33 

•578 

17.0 

.669 

.88 

-381 

5-o 

.197 

3  -°° 

.560 

18.0 

.709 

•33 

•37° 

5-5 

.217 

2-73 

•545 

(From  Hartmann-Knoke,  "Die  Pumpen.") 


At  any  point  of  the  stroke    A»=I  o 

0.9 
•fvlh=ApVi±AcV2,      From  (52) 


dh    d  /AS±A,V2\ 


(54) 


At  the  end  of  the  stroke  v% 
may  be  assumed  constant  and 
small  of  value,  while  at  a  crank 


0.8 
0.7 
0.6 
0.5 

0.4 
0.3 
0.2 
0.1 


01234     6     8     10    12   14    16    18=7i 


position  of  90°  the  valves  may  FlG"  '87-V*™tion  of  jM*h  Vaccord- 

ing  to  Hartman  and  Knoke. 

be   assumed    to  be  wide  open 

and  V2  may  be  taken  as  zero.     At  each  of  these  points  b  may  be 

considered  constant.     These  assumptions  then  give  the  following: 


DYNAMICS  OF  WATER  END  217 


* 

Aru>2( cos  8+-  cos  26 ) 

dt  \  n  ] 


Mo 
at  90 


If  now  the  value  of  bo  be  assumed  and  ho  be  equated  to 
the  size  of  the  valve  or  valves  may  be  found  from  (55).  Eq. 
(56)  may  be  used  to  find  Amax  if  &max  is  assumed  or  by  assuming 
/jtnax  the  value  of  bmax  may  be  found.  Of  course  the  value  of 

ju  depends  on  the  value  -7  and  must  be  assumed  and  then  the 

correct  value  used  for  the  second  approximation.     The  equa- 
tions above  are  reduced  into  other  forms. 

N 
Air—-  =q=  A2T  —  , 

60      "  27r' 


,  A  pTT^CO  4  60  ___  I 

=  =  (0.85  X2  ~  20°     ' 


^T  q  ,     . 

bod==  —   ~     —=^>=o'ii2wN  or  aii24-/^   (57) 

2 


or 


N 

27T 

-  •  •  •  (58) 


In  these  two  formulas  bo  may  be  assumed  (usually  as  i  to  3 
feet)  and  /  or  d  found. 


218  PUMPING  MACHINERY 

If  one  valve  is  used  q  cared  for  by  this  valve  is  found,  but 
if  nf  valves  are  used  each  valve  is  designed  for  —t  cubic  feet  of 

water  per  second. 

The  Eq.  (56)  becomes     , 

7  QT  Q  /  \ 

/*max=         »— ==— — ^== (59) 


In  this  /^max  is  assumed  and  then  from  — —   the  value  of 

n  is  found  in  Fig.  187  and  then  Z>max  is  found. 

The  force  exerted  by  a  spring  is  proportional  to  the  amount 
of  compression  and  in  general  if  yo  is  the  amount  of  compression 
the  force  F3  is  given  by 


when  C  is  the  force  required  to  compress  the  spring  i  foot. 

If  yo  is  the  amount  of  initial  compression  in  the  spring  in 
feet  the  values  of  b0  and  6max  are  used  in  the  following  equations: 


(6o) 


_ 
-  ;;.v.  .  (60 


Solving  for  yo  and  C  there  results 

\        .    ......     (62) 


Z,  f*\ 

«0  —  -4  —    —  no  .....     (63) 


These  are  the  constants  for  the  spring  to  give  the  desired 
pressures,  and  are  used  in  fixing  the  size  of  the  spring  wire, 
number  of  turns  and  other  dimensions. 

The  formulae  just  used  were  for  definite  velocities,  but  actual 
velocities  are  variable.  This  will  change  certain  of  the  forms, 
as  observed  before  in  the  discussion  of  the  resistance  to  the 
passage  of  the  water, 


DYNAMICS  OF  WATER  END  219 

The  movement  of  the  valve  and  its  change  in  velocity  intro- 
duces a  change  in  the  term  b,  which  has  been  used  to  denote 

W,+F, 

Avw 

On  account  of  friction  of  the  valve  the  force  to  lift  is  given  by 
P  =  W,+FS±R, (64) 

where  R  =  resistance  due  to  friction. 

By  properly  designing  the  valve  and  its  connections  the 
term  R  may  be  made  small.  Hence,  although  not  strictly 
correct,  the  resistance  to  valve  motion  may  be  considered  as 
Wv+Fs  or  when  measured  in  feet  of  water  on  area  of  valve, 
W,+F. 
Avw 

Valve  Pressure  for  Stroke.  The  pressure  loss  for  the  valve 
varied  from  kn  or  h^  Eqs.  (36)  and  (33),  at  the  beginning  of 
the  stroke  to  fe,  Eq.  (46),  during  the  major  part  of  the  stroke 
as  shown  in  Fig.  186. 

Simple  Methods  of  Design  of  Size  and  Number  of  Valves. 
The  method  used  above  for  the  design  of  valves  is  based  on 
theory.  Two  practical  rules  for  the  same  purpose  used  in 
the  United  States  are  those  due  to  Hague  and  to  Reynolds. 
Mr.  C.  A.  Hague  recommends  using  50  per  cent  of  the  piston 
area  for  the  valve  area  on  discharge  or  suction  side  when  the 
piston  speed  is  100  feet  per  minute  and  150  per  cent  when  the 
piston  speed  is  300  feet  per  minute.  For  intermediate  piston 
speeds  a  proportional  value  of  the  percentage  is  used.  Mr.  I. 
H.  Reynolds  uses  a  very  simple  rule:  i  square  foot  of  valve 
area  is  required  on  suction  and  on  discharge  for  each  million 
gallons  capacity.  In  this  way  the  piston  speed  need  not  be 
considered,  as  it  is  the  quantity  of  water  in  a  given  time  which 
fixes  the  area  of  the  valves. 

The  springs  in  this  case  are  not  designed  but  sizes  and  diam- 
eter are  assumed  as  shown  on  page  317. 

All  points  have  now  been  covered  in  the  design  of  valves 
and  the  value  of  £  for  the  valves  may  be  found  and  the  results 
computed.  The  general  equation  for  hs  may  be  formed. 


220  PVMPING  MACHINERY 

The  variation  in  pressure  of  the  water  on  the  bottom  of 
the  piston  on  the  up  stroke  may  now  be  written: 


The  quantities  hno  or  hdvo  on  discharge  are  the  resistances  at 
the  time  of  opening  the  valves.  Their  values  may  be  quite 
large,  and  in  order  to  reduce  them  and  the  acceleration  term 
air  chambers  are  used,  as  will  be  seen  later.  It  is  well  to 
remember  that  hs,  or  the  pressure  under  the  piston  during  the 
suction  stroke,  can  never  be  less  than  zero.  This  means  that 
the  sum1  of  the  resistances  and  the  lift  on  the  suction  side  must 
be  less  than  34  feet.  In  many  cases  pumps  are  run  so  fast  that 
this  quantity  is  exceeded  causing  the  water  column  to  break, 
and  pounding  results. 

In  designing  pumps  with  clack  valves  the  same  methods 
as  used  with  the  lift  valves  may  be  employed,  remembering 
that  moments  of  forces  about  the  pivot  must  be  considered 
and  also  that  the  perimeter  of  discharge  is  not  the  complete 
perimeter  of  the  valve.  The  lift  h  may  be  taken  as  the  lift 
at  the  outer  end. 

Problem  in  Valve  Size  and  Pressures.  A  problem  will  be 
computed  to  illustrate  the  use  of  the  above  formulae: 

Suppose  a  20  by  30-inch  duplex  single-acting  pump  is  driven 
at  60  R.P.M.  and  that  there  is  9  feet  of  1 8-inch  suction  pipe 
with  one  bend  followed  by  i  foot  lift  at  pump  to  deck  of  valves, 
and  300  feet  of  1 8-inch  discharge  pipe,  rising  200  feet  and  con- 
taining three  right-angle  bends  of  6-foot  radius.  The  number 
of  valves  and  the  variation  of  suction  pressure  will  be  found. 

Losses  in  Pipe  Lines. 

Area  of  suction  and  discharge  pipes: 

^(i.S)*-i.77. 

4 


DYNAMICS  OF  WATER  END  221 

Area  of  water  piston 

4     \12/ 
Loss  at  entrance 

2g     \i-777 
In  suction  pipe 

d  2g 
For  dirty  pipes 

/-0.03; 


._Q 

\I-77/    2g  2g 


One  bend 

£  =  6  feet; 


k}~d  =  2.  15X0.03X^=0.41; 

/2.i8\252        ,  52 
^  =  0.41  (-    -    —  =  0.01  —  . 

\1.77/    2g  2g 

Sudden  enlargement  into  suction  chamber, 

/      AA2vs2     [      i.77\2/2.i8\252  S2 

h=    i—  —     —  =    i  --          "   '     —  =  0-0—  • 


-05 

A    /     2g       \          2.l6/     \1.77/     2g  2g 

The  total  term  will  then  be 

(A        \S2  S2          S2 

2w-  —  hi    —  =  [l.  5+0.27-4-0.61  +0.05  +  1]—  =3.40—  . 
Al         )2g  2g  2g 

The  terms  show  the  importance  of  the  various  friction  losses. 
In  the  present  instance  it  is  seen  that  the  loss  at  entrance  is 
the  most  important  item. 

On  the  discharge  side  the  equivalents  of  these  terms  are 
given  below: 


222  PUMPING  MACHINERY 

Loss  in  friction  of  pipe 

=  0.03X300/2.  i8\252 
i-5       \«-77/  2g 

S2 
=  9. 

Loss  in  bends 


Loss  due  to  sudden  contraction 


Final  velocity  head 


S2 

0.04  —  . 


i.77 
S2 


TzrVsT  °-3« 

The  next  term  of  the  expression  for  h,  is  that  involving 
the  acceleration.    Let 


Ai      1.77 
and 


^^  j 

This  total  term  may  be  divided  into  two  parts,  one  —   and 

o 


the  other  +  ^-    The  latter  equals  ^a  =  .37a. 

AI         A2  >g  32-2 


DYNAMICS  OF  WATER  END  223 

The  value  of  xa  is  found  by  multiplying  the  expression  for 
X  and  a  in  Eqs.  (2)  and  (4)  or  by  multiplying  the  values  of  x, 
and  a  if  the  numerical  values  of  x  and  a  at  different  angles  are 
found  as  shown  on  page  207  in  terms  of  r2u2. 

These  values  may  be  plotted  for  different  piston  positions. 

Size  and  Number  of  Valves.  The  American  method  of 
designing  pumps  is  to  use  a  large  number  of  small  valves  in 
place  of  a  small  number  of  large  valves. 

The  reason  for  this  is  the  fact  that  since  the  area  of  a  valve 
varies  as  the  square  of  its  diameter,  and  its  discharge  area 
varies  with  the  first  power  only,  smaller  valves  have  a  larger 
discharging  area  for  the  same  area  of  valve  deck.  To  util'ze 
the  central  part  of  large  valves,  they  are  made  multiported, 
as  shown  in  Fig.  188.  It  is  this  type  of  valve  which  will  be 
compared  with  the  small  American  valves  For  one  pump 
cylinder: 

AN     30X314.  i6v  60  ,  .    ,. 

q  =  2rA—  =  -  -    ~  —  X  —  =  5.5  cubic  feet  per  sec. 

DO  1728  OO 

0.1124—  ,N 
60=-~-; 

assuming 


2  ft. 

0.1124X5.5X60 
-  - 


=  57  valves; 


fo  =  -!-</  =  —  =0.00167  ft. 

200          200 

To  compare  the  results  for  the  4-inch  valves  with  another 
method  used  to  determine  the  number  of  valves,  consider  the 
rule  that  the  velocity  through  the  area  occupied  by  the  valves 
should  be  between  100  and  125  feet  per  minute.  Since  the 
water  is  delivered  through  the  valves  in  one  stroke  or  a  half 
revolution,  this  gives 

29X60     ^2X5.5X60 
~  100  X  (area)"  '_4^       '^' 

100  X  144 


224 


PUMPING  MACHINERY 


FIG.  188. — Multiported  Valve. 


American  practice  is  to  use  4-inch  valve,  and  this  diameter 
will  be  assumed. 


DYNAMICS  OF  WATER  END  225 

This  means  about  200  to  250  feet  per  minute  through  the 
opening  in  the  valve  deck.  If  this  is  figured  on  the  maximum 
velocity  through  the  opening  in  place  of  the  mean  this  number 

75  will  have  to  be  increased  by  the  ratio  ofir:2or^  =  -X75  =  n8. 

This  result  is  different  from  the  previous  result.     Using  the 
Hague  and  Reynolds  methods  the  following  results: 

Piston  speed  2LN  =  2  X  if  X  60  =  300. 

Hague's  method: 

Area  piston  =2.18  sq.  ft.; 

Valve  area  =1.50X2.18  -=-3.  25  sq.  ft.; 

Number  of  valves  =  —  —  =37  valves. 

49 
Reynolds'  method: 

5.5X60X60X24X1728 
Capacity  of  f  pump  -  =3,570,000 

gallons  per  24  hours. 
Area  =  3.  5  7  sq.  ft. 

Number  of  valves  -^  =42  valves. 

7T  I 

49 

These  two  results  agree  fairly  well,  but  they  differ  from  the 
value  57  first  obtained.  The  reason  for  this  difference  is  the 
fact  that  bo  =  2  feet  is  smaller  than  the  amount  used  in  either  of 
these  latter  methods. 

For  the  multiported  valve  several  formulae  will  be  derived 
referring  to  Fig.  188: 


for  each 


(66) 
(67) 


226  PUMPING  MACHINERY 

Hence 


A,    ?A    irbiZd    bi 

~~-;    .....     (68) 


i  A,      bi 

hQ=  --  -  =  —  . 
50  /      100 

From  Fig.  184 


Now  W  =  - 

Assuming,  as  before,       bo  =  2  ft, 


.. 

27T 

From  the  figure 

-(d1-x)2+ird1X+Trd2X+   .   .   .    =-[ 
4  4 

or 

—  irnx2  +  wn 


n2-n 
Assuming  d\  =  6  in.  and  n  =  5, 


0       0 

Let  ^1=0.  19  ft; 

62=0.15  ft; 
d0ut=di+(2n-  1)^  =  ^+9X0.34  =  3.  56; 

,  I    ,          O.IO 

ho  =  —  bi=  —  —  =  0.0019. 
100         100 


DYNAMICS  OF  WATER  END  227 

Comparing  results,  assuming  50  per  cent  as  the  excess  area 
required  for  valve  deck,  it  is  seen  that  the  57  valves  require 
745  square  feet  of  area  while  the  multiported  valve  requires 
12.7  square  feet. 

This  calculation  shows  the  necessity  of  a  large  number  of 
valves  when  a  high-speed  pump  is  used.  If  the  speed  of  the 
pump  is  decreased  the  number  of  valves  will  be  diminished  in 
proportion  to  the  square  of  the  speed,  as  the  number  varies 
as  the  product  qN,  in  which  q  depends  on  N.  Thus  for  a  20- 
revolution  pump  the  number  of  valves  would  be  decreased 
to  iXii2  =  i2,  and  the  diameter  of  the  multiported  valve 
would  be  changed  to  7X3.35,  or  1.12  feet.  The  area  required 
would  be  0.83  square  foot  and  1.94  square  feet,  respectively, 
for  the  small  valves  and  the  multiported  valve.  By  changing 
bo  the  number  and  size  could  be  diminished.  This,  with  a 
3-foot  pressure  for  bQ  in  the  original  calculation,  results  in  the 
following: 

n  =  $S. 

Jo  =  2.  5  approx. 

Area  valve  deck  =4.96  sq.  ft. 

Area  multiported  valve  =  3:52  sq.  ft. 

To  find  the  spring  pressure  at  the  position  ho  use  the  follow- 
ing formula: 


For  small  valve 

6i.5X2X^ 
144 

assuming  Wv  =  %lbj 

10.90=1+^0; 

.'.   FQ  =  10.40. 
For  large  valves 

62.5X2X^X0.19X9.3  = 

6Q3= 
Let 

PFP  =  9ol 

/.  ^0  =  603. 


228  PUMPING  MACHINERY 

Now 


Since  hmsix  will  be  greater  than  ho,  consider  it  as  0.035  f 
the  first  approximation  for  small  valves. 

-  -S"        -  =  ^  =  2.4; 

M=-53; 

6m=WV^J  =  \57Xo.53X8.o2X.33X.o35J  =  3 -87  ft. 
For  the  large  valve  let 

hm  =  .02 
so  as  to  give  less  distance  for  this  large  valve  to  seat, 


2km       2X.02 

Hence 

^1  =  0.64; 


_(  5-5      r        V    n,ft 

\.64X8.02X  2  Xg.3X.02/ 


Since 


W,+F0  =  62.  5  X8.4X7TX.  19X9.3  =  2920. 

Spring  Constants.     In  the  case  of  the  small  valve,  the  spring 
tension  which  was  10.40  Ibs.  at  .0017  movement  will  be  20.6  Ibs. 

20.6  —  10.40 
at  .035  movement,  or  •/  -  —  -  x  —  =26  Ib.  per  inch  of  com- 

pression (approx.). 

For  the  large  valve  this  becomes 

(28^0)—  603  2227 

•T^-  ON     =—        '       =9660  Ib.  per  inch,  approx. 

(.02  —  .0008)12     .0192X12 


DYNAMICS  OF  WATER  END  229 

In  this  case  there  would  be  several  springs  to  hold  such  a 
large  valve  to  its  seat,  say  ten.  This  would  give  966  Ib.  per 
inch  for  the  modulus  of  these  springs.  The  large  valve  as 
designed  would  be  very  cumbersome  and  would  undoubtedly 
be  divided  among  a  number  of  smaller  valves,  changing  the 
results  considerably.  It  is  suggested  that  the  student  work 
out  the  data,  using  six  valves  of  the  same  style  in  place  of  one 
large  valve. 

The  method  of  designing  the  springs  will  be  considered 
in  the  next  chapter. 

Valve  Resistance.  The  resistance  of  the  4-inch  valves  during 
the  operation  of  the  pump  is 


For  disc  valves  the  diameter  of  the  passage  is  4  —  2X1^  =  3.2 
inches  or  0.27  foot  and  from  this  the  following  results: 

;-wr 


=  .55+0.00+8.9  =  9.4 

1.18 


64.32X3.28       I2 


=3.8  ft. 
° 


The  value  for  hv  is  the  resistance  during  the  stroke  of  the 
pump.  It  should  be  slightly  larger  than  the  valve  bmeai  because 
it  includes  the  losses  in  the  passages.  £max  was  taken  as  3  feet 
in  working  out  the  spring  constants. 

The  resistance  to  water  passage  through  both  tH  suction 
and  discharge  valves  is  the  same.  The  resistance,  however,  at 
each  end  of  the  stroke  is  different  for  these,  as  at  that  time 
inertia  plays  an  important  part,  and  this  effect  depends  on  the 


230  PUMPING   MACHINERY 

length  of  water  column,  which  is  greater  on  the  discharge  side. 
The  equations 

i\A.-A.    W.F.     W,a  A 


and 


give  the  loss  of  pressure  at  opening  of  the  valves,  and  to  get 
numerical  values  for  these  the  pumps  with  the  4-inch  valves 
will  be  considered,  assuming  that  the  suction  lift  is  equivalent 
to  10  feet,  including  friction;  that  the  discharge  head  is  equiva- 
lent to  232  feet  with  friction;  that  there  are  10  feet  of  i8-'nch 
suction  pipe  and  300  feet  of  1 8-inch  discharge  pipe,  and  that 
the  4-inch  valves  have  3.2-inch  openings  beneath  them. 
At  the  end  of  the  stroke 


a  = 


=  58  f  *'  per  sec'  per  sec' 


57^.272  =  3.  28  sq.  ft; 


A  =  2.18  sq.'ft.; 


==^     7  =i-77  sq.ft. 
4  J44 


From  Eq.  (31  a) 


f 
-- 


=  934X62.5  =  58200  Ib.  per  sq.  ft. 


The  effect  of  inertia  in  the  pipe  line  is  seen  to  be  much 
greater  than  the  static  head.  Of  course  this  water  probably 
does  not  start  from  rest  each  time  and  hence  such  a  high  value 


DYNAMICS  OF  WATER  END  231 

of  pud  may  not  occur.     If  this  does  not  occur  the  water  column 
must  separate  and  pounding  will  be  observed. 

7-3.28     57X0.5     10.40X57 
_ 


+57X^5      ^8oft 
32.16  3     3-282J 

This  large  resistance  at  opening  of  the  valve  is  due  to  the 
fact  that  a  large  discharge  pipe  is  used  with  no  air  chamber; 
the  effect  of  such  a  chamber  will  be  seen  later. 

pls=    24—10  — — — 58   62.5  =  2X62.5  =  125.0  Ib.  per  sq.  ft. 

The  result  shows  that  the  net  pressure  causing  flow  in  the 
suction  pipe  is  only  2  feet,  and  hence  a  slight  increase  in  the 
speed  of  the  pump  would  cause  the  suction  column  to  break. 

,          i    [      4-97-3-28  ,57X0.5     10.40X57 

Pn  —  y  I    I  2  S 

62.51  4-97  4-97  4-97 

57X0.5  r       2.18       1 

+— 7^587 TT — ^    =2.7  ft.,  approximately. 

32.16  3  (4.97X3.28)] 

Plotting  Pressures  for  Points  of  Stroke.  The  expressions  for 
all  of  the  terms  of  Eq.  (65)  are  known  at  different  positions  of  the 
pistons.  The  terms  are  the  following: 

1 .  The  curve  of  friction  varying  as  S2  known  as  the  KS2  curve. 

2.  The  piston  movement  curve  or  the  x  sin  $  curve.     (This 
is  zero  for  horizontal  pumps.) 

3.  The  inertia  curves;  one  for  th?  a  term  or  Ka  curve  and  the 
other  for  the  xa  term  or  the  Kxa  curve. 

4.  The  valve  curve. 

These  values  of  the  different  terms  of  Eq.  (65)  are  now 
plotted  in  Fig.  189,  in  which  the  resultant  pressures  above 
the  absolute  zero  are  given  when  0  =  90°. 

It  is  to  be  remembered  in  this  work  that  dimensions  are 
in  feet. 

In  Fig.  189  the  various  curves  are  plotted  for  different 
piston  positions.  The  positions  of  the  piston  for  the  crank  posi- 
tions of  o°,  30°,  60°,  90°,  120°,  150°,  and  180°  have  been 


232 


PUMPING   MACHINERY 


shown  by  vertical  lines.  The  important  lines  of  pressure  are 
marked  in  different  ways.  The  atmospheric  pressure  of  zero 
gauge  pressure,  or  34  feet  absolute  is  shown  by  a  dot-and-dash 
line  while  the  lift  h2  of  10  feet  is  represented  by  a  dash-and- 
three-dot  line. 

The  friction  loss  varies  with  S2  and  is  given  by  a  dash-and- 
two-dot  line.  This  curve  starts  from  zero  at  the  piston  posi- 
tion for  o°  and  reaches  a  maximum  just  before  90°,  while  the 
value  at  180°  is  again  zero. 

The  valve  friction  is  represented  by  a  dash  and  two-dot 
line  and  remains  practically  constant  during  the  larger  por- 


-34G 


Atmospheric  Pressuro  34 'Absolute 


FIG.  189. — Pressure  on  Suction  without  Air  Chamber. 


tion  of  the  stroke  according  to  the  experiments  of  Hartmann- 
Knoke. 

The  portion  of  the  inertia  effect  due  to  the  water  in  the 
cylinder  has  been  shown  to  vary  with  the  product  xa.  This 
is  represented  by  a  dash  line  which  starts  at  zero,  increases  to 
a  maximum  about  45°,  passes  through  zero  just  before '90° 
and  increases  negatively  to  the  end  of  the  stroke.  This 
quantity  is  the  same  for  each  end  of  the  cylinder  and  for 
each  stroke. 

The  chief  disturbing  element  in  the  suction  pressure  is  the 


DYNAMICS  OF  WATER  END  233 

inertia  of  the  water  in  the  pipe  line.  This  is  the  quantity 
given  by  Y  =  ka  and  is  shown  by  the  dotted  line.  The  water 
in  the  suction  pipe  has  to  have  an  acceleration  proportional  to 
that  of  the  piston  if  the  column  remains  intact,  and  conse- 
quently to  produce  this  a  great  head  results. 

The  curve  representing  the  head  lost  due  to  piston  travel, 
x,  is  shown  by  a  slanting  dot-and-dash  line. 

The  combined  curve  is  drawn  as  a  solid  line.  At  the  begin- 
ning of  the  stroke  the  gauge  pressure  is  —34  feet,  or  absolute 
zero.  This  means  that  the  column  would  probably  break  and 
produce  knocking.  The  pressure  increases  rapidly,  however, 
reaching  a  value  above  the  atmosphere  at  the  end  of  the  stroke. 
This  shows  that  although  the  water  is  lifted  10  feet  +  2|-  feet, 
the  inertia  of  the  water  in  the  pipe  is  so  great  that  on  trying  to 
stop  it  the  pressure  increased  to  the  value  shown  in  the  figure. 

The  pressures  marked  on  this  and  the  other  pressure  dia- 
grams, Figs.  191,  197  and  198,  are  measured  in  feet,  starting 
from  atmospheric  pressure.  The  absolute  zero  of  pressure  is 
34  feet  below  this. 

The  absolute  pressure  of  this  figure  at  its  lowest  point  can 
never  be  a  negative  quantity,  as  there  is  never  a  negative  abso- 
lute pressure.  Should  the  diagram  give  this  in  the  combina- 
tion of  the  various  lines  this  would  indicate  that  the  water 
column  would  separate  and  shock  would  result.  This  occurs 
when  the  pressure  assumes  an  apparent  negative  value.  hs  is 
therefore  limited  to  zero.  This  value  cannot  be  obtained,  as 
the  pressure  of  boiling  for  the  water  temperature  would  fix  the 
lowest  value.  When  this  pressure  is  reached  the  column  will 
separate,  since  steam  will  form.  This  gives  then  as  a  limit 


(I  A  \2       \S2 
1  '\~A~J  +  /2£ 


At  the  end  of  the  upper  stroke  when  S  =  o 


i          "~-iou  -A   -#180  /      \ 

-/^ —  -2-JpL ,          (71) 

1         o 


234 


PUMPING  MACHINERY 


or 


L< 


a  so 

g 


.     .     (72) 


at  90' 


L< 


-      •     (73) 


Ai  g 
LIMITS  OF  HEAD  FOR  PRIMING 

These  give  the  limiting  values  of  L  with  which  the  pump 

will  operate  without  breaking 
the  column  of  water  or  pound- 
ing. Should  L  be  made 
greater  the  pump  will  pound. 
At  times  small  pumps  are 
built  with  foot  valves  at  the 
base  of  the  pump  and  a 
bucket  plunger  in  the  same 
barrel  at  some  height  above. 
Such  pumps  have  a  limiting 
distance  between  the  piston 
and  foot  valve.  Consider  the 
pump  (Fig.  190)  to  have  the 
same  bore  as  the  pipe  below, 
and  that  the  water  has  been 
sucked  in  until  it  reaches  the 
height  h,  at  which  time  the 
action  stops  because  when 
the  piston  is  at  its  lower  end 
of  stroke  the  pressure  of  the 
air  is  just  atmospheric,  while 


FIG.  I90.-Suction  Lift. 

That  is,  the  pressure  will  not  be  sufficiently  low  to  allow 
any  more  water  to  enter  when  at  the  top  of  its  stroke.  The 
following  equation  holds  then  from  Boyle's  law: 


DYNAMICS  OF    WATER   END  235 

(ha-h)A(s+d-h)=haA(d-h); 

(ha-h)(s+d-h)=ha(d-h); 

has+ha(d-h)-(s+d)h+h*=ha(d-h)i 


(74) 


Now  when  ^(s-}-d)2=has  there  are  two  equal  roots  for  h 
or  the  water  will  rise  to  a  certain  point. 

If  ^(s+d)2>has  there  are  two  real  roots,  which  means  that 
the  water  will  rise  to  a  certain  point,  but  if  it  could  be  moved 
beyond  this  point  it  would  go  still  further  to  another  point  h. 

If  i(s+d)2<has  there  will  be  no  real  roots  or  the  water 
will  be  sucked  up  and  the  pump  will  operate. 

This  gives 


or 

d<2\/h~J-s  .........     (75) 

This  then  gives  the  limiting  lift  for  such  a  pump  when  the 
suction  valve  is  placed  in  the  bottom  of  the  pipe. 

The  same  principle  is  applicable  to  pumps  in  which  the 
volume  of  the  cylinder  above  valves  is  large  compared  with 
the  displacement  of  the  pump.  Let  Vc  be  the  volume  of  the 
cylinder  above  valves  when  the  piston  is  at  the  end  of  the 
stroke  (called  clearance  on  steam  cylinders)  and  Vr  the  volume 
displaced  by  the  piston,  then  if  the  pump  is  to  operate  under 
the  suction  lift  h8  to  the  suction  valves  without  filling  the 
pump  barrel  with  water  to  prime  the  pump  for  starting,  the 
following  equation  must  hold: 


Under  this  condition  the  air  in  the  cylinder  on  the  stroke  at 
which  water  just  reaches  the  suction  valve  will  be  compressed 
to  a  little  over  atmospheric  pressure  and  be  driven  out.  This 
gives 


236  PUMPING   MACHINERY 

Vc  is  really  what  might  be  called  the  clearance  volume,  and 
it  is  seen  that  when  this  is  zero  the  limiting  suction  head  is  7/a; 
when  it  is  equal  to  Vr,  ha  becomes  \ha  at  the  limit. 

The  limitations  on  this  deduction  must  be  kept  in  mind. 
The  value  of  hs  in  formula  (76)  is  the  suction  head  against 
which  a  pump  with  air-tight  valves  will  operate  without  priming. 
If  the  pump  is  primed,  however,  this  water  fills  the  clearance 
space  and  Vc  becomes  equal  to  zero,  and  then  the  pump  will 
operate  against  the  limiting  head  ha  if  no  friction  is  considered. 
If  friction  [is  considered  the  values  given  in  Eqs.  (72)  and  (73) 
are  the  limiting  suction  lengths  used  under  these  conditions. 

Discharge  Pressure  Curves.  Passing  now  to  the  discharge 
side  it  is  seen  that  the  expression  for  the  pressure  below  the 
piston  on  the  down  stroke  will  be 


As  before  the  various  losses  will  be  plotted  in  Fig.  191. 
To  plot  these  losses  for  the  problem  stated  throughout  this 
chapter,  the  following  numerical  values  hold: 

^  =  345 
£3  =  12.6-; 
hf2  =  200; 


7^0  =  480  ft; 
7^  =  3.8. 

Assume  that  all  valves  are  open  wide. 
xa 


L+2Ll=H,=? 

Ai         A2     )g  Vi-77 

These  are  now  plotted  in  Fig.  191. 


DYNAMICS  OF  WATER  END 


237 


This  figure  shows  the  great  effect  of  the  inertia  of  the  water, 
and  in  working  it  out  the  effect  of  the  compression  of  the 
water  has  been  neglected,  as  the  use  of  the  air  chamber  makes 


FIG.  191. — Pressure  on  Discharge  without  Air  Chamber. 

such  a  condition  impossible.  However,  the  problem  has  been 
computed  to  show  what  is  eliminated  by  the  use  of  the  air 
chamber. 

The  scale  of  head  has  been  increased  so  that  the  heads 
could  be  plotted  and  so  great  is  the  scale  that  the  loss,  due  to 

Xd 

friction,  the  larger  part  of  the  valve  friction,  and  the  — =Z 

o 

terms  do  not  show.  The  important  curves  are  the  static 
head,  234  ieet  above  the  absolute  zero,  the  curve  due  to  acceler- 
ation of  the  water  in  the  pipe,  Y=Ka,  and  the  atmospheric 
line  at  34  feet.  These  combine  with  the  initial  valve  friction  to 
give  a  line  which  starts  with  a  value  of  over  800  feet  gauge 
pressure  and  ends  with  a  pressure  below  absolute  zero.  This 
last  pressure  is  impossible,  and  as  a  result  water  will  separate 
from  the  piston  or  water  may  be  sucked  through  from  the 
suction  side. 

The  great  variation  should  be  eliminated,  and  one  method 
is  to  run  the  pump  more  slowly  or  cut  down  the  mass  of  water 
which  has  to  be  accelerated.  This  latter  is  done  by  the  use 
of  the  air  chamber,  as  will  be  shown. 

The  diagrams  seen  here  and  in  Figs.  174,  175,  189,  197, 


238 


PUMPING  MACHINERY 


and  198  have  been  on  time  and  space  bases,  and  it  is  well  to 
see  how  a  curve  on  one  base  may  be  changed  to  one  on  another 
base. 

Fig.  192  gives  a  velocity  diagram  ABC  on  a  time  base. 
To  change  this  to  a  space  base  the  curve  DEF  has  been  drawn 
with  an  integraph  such  that  the  height  i"  i'"  directly  over 
the  point  i  is  proportional  to  the  area  An',  2"  2'"  =  area 
.422',  etc.  The  planimeter  may  be  used  for  the  determination 
of  these  areas  and  from  them  the  ordinates  i"  i'",  2"  2'",  etc. 

Now   area  An'  =  I  vdt=  space.     Hence     the   ordinates   of 
JQ 

the  curve  DEF  are  the  spaces  passed  over  in  the  times  repre- 

Velocity 


Time 
FIG.  192. — Velocity-space  from  Velocity-time  Diagram. 

sented  by  the  points  i"  2",  etc.,  or  the  curve  is  a  time  space 
curve. 

If  now  the  horizontal  lines  i"  iiv,  2"  2iv  be  drawn  and  the 
distances  iiv  iv,  2iv  2V,  be  laid  off  equal  to  i'  i,  2'  2  respec- 
tively, the  curve  FGH  is  obtained  which  is  the  velocity  curve 
on  a  space  base. 

The  curve  in  Fig;  193  is  a  velocity  curve  on  a  space  base. 


DYNAMICS  OF  WATER  END 

Velocity 


239 


Ni    \\   \     \! 
1      C     62 


Space 
FIG.  193. — Velocity-space  Diagram  to  Velocity  Time. 

In  this  the  method  of  getting  to  a  time  base  is  not  direct.    The 
following  method  is  used: 


ds'  =  vdt; 

&=*£• 


(itf. 

J  v 


Hence  the  time  taken  to  go  a  distance  s  may  be  expressed 
as  the  area  of  a  curve  if  the  quantity  —  is  an  ordinate  and  the 

space  passed  over  is  the  abscissa. 

The  method  of  procedure  is  to  construct  a  reciprocal  curve 

to  the  one  given  in  which  the  ordinates  are  quantities  —  and  to 

draw  the  integral  curve  of  this. 

To  get  the  reciprocal  curve  of  BCD  construct  the  line  EF 
at  any  distance  A  from  the  base  and  on  the  ordinate  i,  iv 
desired  swing  la  down  to  b,  join  iv  and  b  and  draw  ac  parallel 
to  ivb,  and  the  distance  ic  will  be  proportional  to  the  reciprocal 


240  PUMPING   MACHINERY 


of  v.    This  distance  is  then  swung  up  to  i1  giving  i,  i1  propor- 

lal  to  - 

v 

This  is  seen  as  follows: 

ic      ib 


tional  to  - 
v 


la     1 


hence  x=A2-. 

.  v  „, 


The  ordinate  of  the  curve  HI]  when  divided  by  A2  gives 
the  reciprocal  of  v.     If  A  be  made  unity 


The  integration  of  the  areas  under  HIJ  gives  the  integral 
curve  KLM  as  before,  and  the  velocity  curve  is  constructed 
by  laying  off  the  distances  i,  iv,  2,  2V  on  the  proper  horizontal 
lines. 

For  purposes  of  studying  forces  acceleration  diagrams  are 
desired.  To  construct  such  from  various  points  of  the  crank 
or  piston  movement  of  a  pump  when  the  velocity  is  known,  the 
following  methods  are  used: 

Let  Fig.  194  represent  the  variation  of  velocity  with  the 
space  passed  over  by  the  body  having  that  velocity. 

In  this  -3-  =  tangent  of  the  angle  of  inclination  of  the  tangent 

to  the  curve  with  the  horizontal. 
Now 

—  \c—=  ——  — 
dsr    dt  ~~      dt~ 


Hence,  if  V  is  multiplied  by  —  ,  the  acceleration  is  obtained. 


DYNAMICS  OF  WATER  END 


241 


FIG.  194. — Acceleration  from  Velocity-space  Diagram. 

V-j-=ab  if  ac  =  V.     The  subnormal  of  the  sv  curve  on  the  s 
ds 

axis  represents  the  acceleration. 

If  V  is  on  a  time  base  (Fig.  195)  the  operation  is  different. 

dv 

In  this  case  -^-=a=the  tangent  of  the  angle  of  inclination  of 
(M>  » 

the  tangent  of  the  curve  with  the  axis  of  t. 


Tfme 


FIG.  195. — Acceleration  from  Velocity-time  Diagram. 

If  then  ad  is  laid  off  from  the  point  of  tangency  equal  to 
unity,  ed  will  equal  the  acceleration. 
Since 

,        , ,  ,dv  dv 

ed  =  adta.n  ead  =  ad-jr  =  iX-Tr=a. 
at  at 

In  the  case  of  the  figures  in  Fig.  175,  showing  the  discharge 
of  the  water  from  pumps  of  various  kinds,  these  are  really  on  a 
time  base  as  they  represent  velocities  of  the  pistons,  or  quanti- 
ties of  water  passing  through  the  pumps  per  second  at  any 
instant,  and  therefore  velocity  in  the  main,  for  different  posi- 
tions of  the  crank  which  is  moving  at  a  uniform  rate. 

The  acceleration  diagrams  could  therefore  be  computed  by 
the  second  method  if  desired. 

In  the  figures  and  in  the  computation  of  the  h8Va  and  hdvo 
the  effect  of  inertia  in  the  suction  pipe  and  discharge  pipe  can 


242 


PUMPING  MACHINERY 


readily  be  seen.  These  masses  of  water  have  to  change  their 
velocity  as  the  piston  changes  its  velocity  or  if  the  pump  is  a 
multicy Under  pump  the  curves  which  have  been  computed  for 
the  resultant  discharge  (Fig.  175)  give  the  velocity  variations 
which  exist. 


FIG.  196. — Pump  with  Air  Chambers. 

To  cut  down  the  variation  of  the  velocity  change  in  the 
suction  and  delivery  mains  and  to  make  the  inertia  forces  smaller, 
air  chambers  are  introduced  on  the  suction  and  discharge  sides 
of  the  pump.  The  object  of  the  air  chamber  is  attained  where 
so  designed  that  the  variation  of  pressure  in  it  will  be  slight 
and  hence  keep  the  water  in  the  main  under  a  practically 


DYNAMICS  OF  WATER  END  243 

constant  velocity  while  the  water  which  is  accelerated  is  only 
that  which  lies  between  the  pump  and  the  air  chamber. 

The  air  chambers  on  the  two  sides  of  the  pump  (Fig.  196) 
are  subject  to  a  certain  amount  of  variation  of  pressure  and  in 
making  the  preliminary  investigation  it  is  well  to  consider 
the  mean  pressure  in  the  air  chamber  as  constant. 

This  pressure  is  that  within  the  suction  air  chamber  which 
causes  a  flow  from  the  forebay  overcoming  the  resistance  of 
entrance,  friction,  and  lift.  There  is  no  velocity  considered 
since  the  head  used  in  causing  the  velocity  in  the  pipe  is  re- 
turned when  the  water  is  I  rought  to  rest  in  the  suction  air 
chamber.  In  the  discharge  air  chamber  the  pressure  is  used 
to  give  the  velocity,  lift  the  water,  and  overcome  friction  of  all 
kinds.  The  pressure  in  the  suction  chamber  is  less  than  at- 
mospheric pressure  in  general,  although  water  might  be  sup- 
plied to  the  pump  under  pressure  from  a  higher  source.  The 
pressure  in  the  discharge  chamber  is  greater  than  that  of  the 
atmosphere.  In  any  case  pressures  are  always  measured  from 
absolute  zero.  The  pressure  in  the  air  chamber  is  measured 
when  there  is  no  velocity  head. 

hsc  =  ha-(^  +  ^~-y^  .    .-.    .    .    .    (80) 

A 

+^4,    .  .  .  .   (81) 


where  the  coefficients  £,  refer  to  the  losses  in  the  pipes  due 
to  friction,  and  the  quantities  y  are  the  lifts. 

These  pressures  in  the  air  chambers  are  now  the  datum 
planes  from  which  to  compute  the  pressure  in  the  cylinder 
beneath  the  plunger  on  the  two  strokes;  the  water  in  the  con- 
necting pipes  and  cylinder  between  the  piston  and  the  air 
vessels  being  the  only  water  which  has  to  be  considered  in  the 
expression  for  inertia  and  friction. 

The  expressions  for  pressure  now  become 


244  PUMPING  MACHINERY 

The  term  L±  is  much  less  than  L,  and  consequently  that 
term  is  smaller.  hsv  does  not  have  such  a  large  value  as  hno, 
although  hsv  is  the  same  over  a  large  part  of  its  range. 

hs  must  be  greater  than  h,-,  and  this  inequality  will  give 
the  limiting  suction  height  or  stroke  for  given  conditions,  as  in 
Eqs.  (72)  and  (73). 

The  computations  below  are  made  for  suction  chambers 
of  the  original  pump  with  the  chamber  placed  two  feet  back 
from  the  entrance  to  pump. 

From  Eq.  (80) 

/  2.l8\2 

8  \(5-°x  - 

—  +0.41)—      I>77 
I-5  /         °4-3 

=  34  —  .92—  8  =  25.08  ft. 
From  Eq.  (21)  the  various  parts  making  up  hs  are 

fee  =  25.08  ft; 
hi  =  2  it.; 
x  sin  (j>  =  X', 
A,,  =  3.  8. 

The  value  hsvo  will  have  to  be  computed,  as  the  pressure 
in  the  air  chamber  will  give  a  new  value  of  pis.  There  will  be 
2  feet  of  water  to  be  accelerated  between  the  chamber  and  the 
pump. 


4-97  4-97  4-97 

57X05 


32.16       3.28X4.97 

^JA  \2A2\  ;  ^  /2.i6\2s2         ^ 

2U--      -    =o.03X—   -   -    —  =  0.06—  ; 
\Ai/  \2g/  i.5\I-77/  2g  2g 


xa     „       ,    . 
—  =  Z  as  before  ; 


(A   T  \a     ,  ,a      2.2 

2-—  L*  -=(1.2  +  1)-=  -  a  =  .o6^a. 
\   A2     /g  g     32.2 


DYNAMICS  OF  \\ATKR  END 


245 


-84'G 


Atmospheric  I  ressure   84/Absolute 


Air  Chamber  Pressure  25.08  Ft.Abs< 


6P 

Lift 


Line  of  Valve  Friction 
Line  of  X1  2       - 
Line  of  -*| 
Line  of  uv —  Z 
Line  of  Y 


90° 


FIG.  197. — Pressure  on  Suction  with  Air  Chamber. 

These  values  are  plotted  in  Fig.  197. 

The  pressure  on  the  discharge  stroke  is  given  by  the  equa- 


tion 


A  \2  92 


The  computations  and  curves  for  the  previous  problem  on 
the  assumption  that  the  air  chamber  is  2  feet  from  the  pump 
are  given  below: 

From  Eq.  (81) 


-+20O 


=  34+4.9+200  =  238.9. 
The  parts  of  Eq.  (83)  now  become 
^  =  238.9; 


x  sn 


246  PUMPING  MACHINERY 


,JA  \2/S2\  2   /2.I6 

SU-—     (—    =0.03  X  —   - 
\Ai/  \2g/  i.\i-77/ 


8  s2 

—  =  .12—  ; 


#-  =  Z  as  before; 
* 

_<4   r  a     2.18         # 
2—  Z,5-  =  --  X  2  --  =  0.0760. 
^2     g     i-77       32-2 

The  value  ^  has  also  to  be  found  when  the  air  chamber 
is  used.     In  this  case 


=  I  239+-^  ^-58  162.5  =  244.4X62.5; 


In  the  two  quantities  hsvo  and  //«/„„  the  effect  of  the  air 
chamber  is  seen  in  the  smaller  value  of  the  last  term  due  to  a 
smaller  value  of  L  of  the  water  to  be  accelerated.  The  effect 
of  the  first  term  is  also  to  be  noted.  This  has  increased  the 
value  of  hno,  while  the  last  term  has  been  changed  so  much 
in  the  expression  for  hd,0  that  the  resultant  head  here  has 
been  made  smaller. 

These  values  afe  plotted  in  Fig.  198. 

In  Fig.  198  the  curves  are  all  marked  and  from  the  descrip- 
tion of  Fig.  189  these  may  be  followed.  The  term  involving 
S2  in  (83)  has  a  negative  coefficient  when  combined  and  con- 
sequently this  curve  is  below  the  line  —34'  gauge  pressure. 
The  effect  of  the  reduction  of  the  mass  of  the  water  to  be  accel- 
erated is  seen  by  the  diminished  height  of  the  curve  Y  =  Ka. 
The  curve  Z  =  Kax  and  the  curve  of  x  are  the  same  as  in  Fig. 
189.  The  value  of  the  valve  friction  at  the  suction  end  cf  the 
stroke  has  been  increased,  although  the  major  portion  of  the  line 
remains  at  the  same  height  as  before. 

By  comparing  Fig.  197  with  Fig.  189  the  great  advantage 
of  the  air  chamber  in  making  the  pressure  more  uniform  and 


DYNAMICS  OF  WATER  END 


247 


in  making  it  possible  to  draw  water  at  the  high  speed  of  the 
pump  is  evident.  The  suction  pressure  in  the  pump  cylinder 
never  falls  to  such  a  low  value  that  the  column  would  break 
when  a  proper  air  chamber  is  used. 


200G 


look 


Atm< 


Air  Chamber 


m 


phere 


-M'GT-  -JL^-rfz^: 


re  239'  Abs. 


20' 


Resultant 


210 


Term  X 


210   / 


FIG.  198.  —  Pressure  on  Discharge  with  Air  Chamber. 

Fig.  198  shows  the  action  of  the  air  chamber  in  reducing 
the  Y  effect  and  the  friction  loss  at  the  opening  of  the  valve 
at  the  beginning  of  the  stroke.  All  of  the  curves  of  the  various 
terms  are  shown  except  that  of  the  negative  term  involvingS2. 
This  is  so  small  that  it  would  not  be  appreciable  on  the  scale 
of  this  figure.  By  comparing  this  with  Fig.  191  the  great 
value  of  the  chamber  may  be  seen. 

In  Eqs.  (80)  and  (82)  the  following  conditions  must  hold: 

hsc  >  hsct, 


The  first  of  these  equations  may  be  used  to  find  the  length 


248  PUMPING  MACHINERY 

3'3,   and   the  second   to  get  the  limiting  -length  h±+yz,  or  the 
limiting  speed  2LN  of  the  pump  piston. 

The  second  equation  will  give  the  limiting  length  to  the 
cylinder  if  the  expression  for  hsc  from  Eq.  (80)  is  inserted  in 
the  expression  for  hs  in  Eq.  (82). 

SIZE  OF  THE  AIR  CHAMBER 

The  suction  air  chamber  may  be  considered  to  receive 
water  at  a  steady  rate  Asvs  per  second,  while  it  gives  to  the 
pump  an  amount  which  varies  for  different  crank  positions 
as  shown  by  the  curves  of  Fig.  175.  The  areas  of  these  figures 
represent  quantity  handled  by  the  pump  so  that  the  area 
above  the  mean  curve  represents  the  amount  supplied  by  the 
suction  air  chamber  during  a  certain  part  of  the  stroke  while 
the  area  below  the  mean  curve  represents  the  amount  sent  into 
the  air  chamber  when  the  pump  does  not  require  it.  This 
statement  is  reversed  in  the  discharge  or  pressure  air  chamber 
when  uniform  discharge  is  assumed  AdVd  in  the  discharge 
pipe.  In  this  case  the  area  above  the  line  represents  the  amount 
given  to  the  air  chamber  while  that  below  the  line  represents 
the  amount  given  up  by  the  chamber  when  the  pump  is  giving 
out  less  than  the  quantity  discharged  through  the  main. 

Calling  the  amount  of  water  in  the  air  chamber  at  a,  Q,  the 

excess  areas  abcve  mean  line  ei,  e2,  etc.,  and  the  deficiencies  d\. 

dz,  etc.,  then  the  quantities  at  the  various  points  b,  c,  d,  etc.,  are 

At  (a)  Q, 


"  (c) 
"  (e) 


.....     (84) 


The  upper  sign  would  be  used  for  the  discharge  air  chamber 
and  the  lower  sign  for  the  suction  air  chamber. 

The  difference  between  the  greatest  and  least  of  these 
quantities  represents  the  greatest  variation  of  volume  of  the 
water  in  the  chamber.  Call  the  difference  F. 


DYNAMICS  OF  WATER  END  249 

The  volume  of  the  air  varies  from  a  maximum  to  a  minimum 
as  this  water  comes  from  the  pump  or  suction,  so  that 

iooFmax~Fmin  =  100^  *  the  percentage 

variation  in  volume.  Since  this  compression  of  air  in  the 
chamber  is  isothermal,  pV =k.  The  pressure  maybe  expressed 
in  any  units.  So  the  formula  will  hold  for  the  pressure  ex- 
pressed in  feet  head,  hence 

^sc  '  sc  ==  ™max  '  mm  ~  "min  '  max* 

since  the  maximum  pressure  occurs  with  the  minimum  volume 
and  vice  versa.  Solving  for  Fmax  and  Fmin  in  tetrms  of  V8C, 
and  substituting  in  the  equation  for  the  variation  of  volume 
there  results 


*c  V 

SC          -  '    8C 


%  variation  =  AV  = 


"max  ~" 


8C  ^max       ^min ' 

If  ^max^min  is  taken  as  h2sc  which  makes  h8c  a  geometric 
mean  or  mean  proportional  between  Amax  and  Amin,  the  value 
of  AV  becomes 

AyJ<  max-Amin=J^ 


or  the  percentage  variation  of  pressure  is  the  same  as  the  per- 
centage variation  in  volume.     Hence 


or 

F8C=J.      :.;:;..    (86) 

If  then  Ah  or  the  permissible  variation  in  pressure  is  assumed, 
the  volume  of  the  air  chamber  to  give  this  result  is  known  for  a 
given  F.  The  equation  shows  clearly  that  the  air  chamber  will 
vary  with  F  and  that  whenever  the  resultant  piston  discharge 
approaches  a  mean  line,  the  volume  of  the  air  chamber  decreases. 


250  PUMPING  MACHINERY 

If  the  variation  Ah  is  to  be  made  small,  the  volume  V8C  will 
be  made  large.  It  is  to  be  noted  that  Ah  is  a  ratio  of  variation, 
and  hence  if  the  same  ratio  of  variation  is  assumed,  the  volume 
of  the  air  chamber  for  a  pump  will  be  the  same  for  all  pressures 
or  suction  lifts.  Moreover,  according  to  this  the  air  chamber 
on  the  suction  side  is  to  be  the  same  as  that  on  the  discharge 
side  if  the  same  percentage  variation  is  used,  the  actual 
variation  being  much  less  on  the  suction  side.  This  is  as  it  should 
be,  for  the  pressure  change  on  the  suction  side  is  most  impor- 
tant, as  there  is  only  the  atmospheric  pressure  to  do  the  driving. 
The  size  of  the  complete  air  chamber  should  be  such  that 
when  the  pump  is  shut  down  the  increase  of  volume  due  to 
the  decrease  of  pressure  from  the  elimination  of  friction,  since 
the  water  is  at  rest,  will  not  be  sufficient  to  cause  the  water 
to  be  entirely,  driven  from  the  air  chamber.  The  value  V&, 
computed  previously,  is  the  volume  of  the  air  in  the  chamber 
when  it  is  under  the  pressure  of  operation.  Calling  V'dc  the 
volume  of  the  air  .when  the  pump  is  at  rest 


Where  hd  is  the  head  when  the  pump  is  stationary,  and  h^ 
is  that  while  the  pump  is  operating.  This  volume  F'^  is  then 
the  minimum  volume  to  be  given  to  the  air  chamber  so 
that  it  will  never  lose  the  air  when  the  pump  is  brought  to 
rest. 

Although  this  simple  method  could  be  used,  Hartmann  and 
Knoke  consider  the  discharge  air  chamber  from  the  stand- 
point of  supply  from  the  pump  when  the  pump  is  started 
up. 

Considering  the  pipe  leading  to  the  air  chamber  to  be  of 
area  Ac,  in  which  the  velocity  is  assumed  to  be  Cc,  it  may  be 
said  that  the  quantity  entering  the  air  chamber  per  second 
is  ACCC  cubic  feet.  The  amount  of  water  leaving  the  air  chamber 
per  second  is  AdCa  cubic  feet,  where  Ad  is  the  area  of  the  dis- 
charge pipe  and  C  the  mean  velocity  at  any  instant  in  the 
cross-section. 


DYNAMICS  OF  WATER  END  251 

If  now  Cc  be  taken  as  constant  the  net  amount  of  water 
entering  the  air  chamber  in  any  time  t  after  the  pump  has 
started  is 

q=AcCj-  CAdCddt.      .     .     .     .     .     (88) 

J  o 

It  is  assumed  that  it  takes  this  time  to  get  the  pump  into  uni- 
form motion  and  at  the  end  of  the  time  that  the  air  chamber 
is  giving  back  as  much  water  as  it  receives. 

The  air  in  the  air  chamber  is  originally  under  the  pressure 
hds=ha+hd  in  feet  of  water  where  ha  is  the  atmospheric  pressure, 
and  hd  is  the  vertical  distance  from  the  end  of  the  pipe  to  the 
water  level  in  the  air  chamber.  As  this  quantity  q  enters, 
the  total  pressure  becomes  greater,  say  hd2  due  to  compression, 
so  far  as  the  chamber  is  concerned,  but  produced  in  reality 
by  the  inertia  and  friction  of  the  water  in  the  discharge  pipe. 
Then  if  V ' ^  is  the  volume  of  the  air  chamber 


=  (Vdc-q)hd2,       .....     (90) 
because  the  temperature  of  the  air  may  be  assumed  constant. 


-, 

—  *  dc 


if  hdz  is  considered  variable. 
Now  from  (88), 

dq=AcCcdt-AdCddt. 
Hence 


(91) 


The  pressure  in  the  air  chamber  at  a  certain  instant  is  A, 
while  that  at  the  entrance  to  the  air  chamber  due  to  the  static 
head  is  h0,  and  there  is  an  unbalanced  pressure  of  Ad(h-hds)w, 
which  may  be  utilized. 


252  fVMPlNG  MACHINERY 

This  acts  on  the  water  in  the  pipe  line  Id,  of  weight 

-rr 
at 


and   produces   the   acceleration   -rr.     The   force  of  inertia  is 


—IdAjuv—T-  expressed  in  pounds. 
Hence 


or 


_  Id    dcd 
ctt  =      "^      j  , 
gh-kj 


dh 
V  dc">ds-', 


,  \  r       A  ^  ^J 
--(AcCc-AdCd)dc  = 

/,  ~\hd 
h+^l     •    .    .     (92) 

n  Jhdo 

where  C  =  Cd  at  the  time  h=hd. 

AdCd=AcCc  because  at  the  instant  when  the  greatest  com- 
pression occurs  in  the  air  chamber  there  is  no  net  flow  into 
the  chamber, 

rd-AcCc 

Ld~  A,  ' 
Then 


g       2Ad 

or 


^c  =  -P^ 


r       hd      hdo 


To  illustrate  the  method  used  in  the  design  of  air  chambers, 


DYNAMICS  OF  WATER  END  253 

the  size  of  chamber  for  the  pump  will  be  computed.  The 
pump  will  be  assumed  to  be  double  acting,  and  of  the  dimen- 
sions given  before.  The  curves  of  Case  2,  Fig.  175,  represent 
the  action  of  this  pump,  and  by  using  a  planimeter  on  the 
original  drawing  the  right-hand  excess  area  was  found  to  be 
0.42  square  inch,  the  left  0.39,  the  right-  and  left-end  deficiency 
areas  which  go  together  0.38  square  inch,  and  the  middle 
deficiency  0.43.  This  gives  the  following  variation  in  the 
quantity  of  water  in  the  air  chamber: 

At  i,  v, 

"    2,   0+0.39; 

"  3,  v  +  0.039 -0.43=0 -0.04; 
"  4,  0-0.04  +  0.42=0  +  0.38; 
"  i,  0+0.38  -0.38=0. 

The  greatest  variation  is  from  0—0.04  to  ^  +  °-39»  or  °43 
square  inch  of  diagram  area. 

The  figure  was  originally  drawn  6  inches  long  and  five-eighths 
inch  for  the  mean  height.  The  length  represented  the  time  of 
one  revolution  or  the  angle  turned  through,  and  the  quantity 
discharged  bv  the  pump  was  10.85  cubic  feet  per  second.  The 
scales  of  the  figure  are  therefore  one-sixth  second  per  inch  of 
length  and  17.38  cubic  feet  per  second  per  inch  of  height.  The 
area  scale  is  therefore  2.89  cubic  feet  per  square  inch. 

The  change  in  the  volume  of  water  in  the  air  chamber  is 
therefore 

0.43  X  2.89  =  1.24  cu.ft. 

The  variation  of  pressure  Ah  is  now  assumed  to  be  5  per 
cent  and  the  volume  of  the  air  in  the  chamber  is,  by  Eq.  (86), 

1.24 
v*c  =  —  =  25  cu.ft. 

To  permit  the  pump  to  be  shut  down  the  air  in  the  discharge 
chamber  on  expanding  from  256  feet  pressure  (the  pressure 
in  the  discharge  chamber  during  action)  to  222  feet  pressure 
(static  head),  the  volume  of  the  air  vessel  should  be  such  that 
this  air  is  not  driven  out.  By  Eq.  (87) 


254  PUMPING  MACHINERY 


The  volume  of  the  cylinder  is  5.5  cubic  feet,  and  the  volume 
of  the  air  chamber  above  is  about  five  times  this.  The  vessel 
would  be  30  inches  diameter  by  48  inches  length.  The  great 
size  of  this  is  due  to  the  kind  of  pump  [Term  F  of  Eq.  (86)], 
and  the  allowable  variation  in  pressure. 

The  method  of  design  sometimes  employed  is  to  assume 
the  ratio  of  chamber  volume  to  cylinder  volume,  and  use  this 
only.  The  ratios  suggested  are:  3  and  6  times  the  cylinder 
volume  for  single-acting  pumps,  and  one-half  to  two-  thirds  of 
this  for  double-acting  pumps. 

From  the  above  formulae  the  following  steps  may  be  taken 
to  determine  the  leading  dimension  of  the  water  system  when 
a  given  quantity  of  water,  Q  per  second,  is  to  be  pumped: 

SIZE  OF  PIPES  AND  AIR  CHAMBERS 

The  suction  pipe  should  be  large  and  as  free  from  bends  as 
possible.  In  the  first  approximation  a  velocity  of  3  feet  per 
second  may  be  assumed  for  the  velocity  in  the  suction  pipe. 

Then  A8=—  . 


The  size  of  the  suction  air  chamber  is  given  by  Eqs.  (86) 
and  (87). 

The  discharge  pipe  is  in  many  cases  so  long  that  it  will 
pay  to  compute  several  sizes.  Suppose  a  pipe  is  found  in 
which  the  lost  head  due  to  friction  is  h/  feet  and  by  using  a 
larger  pipe  this  may  be  reduced  to  h'f  feet.  The  gain  by  this 
enlargement  is 

(hf-h'f}wQ 


550  X  eft. 


=  H.P. 


If  now  the  power  costs  M  dollars  per  horse  power  hour 
the  saving  per  year  of  T  hours  will  be 


DYNAMICS  OF  WATER  END  255 

This  same  could  be  used  to  pay  the  difference  in  yearly 
cost  between  the  cost  of  the  small  and  large  pipes.  The 
increased  cost  of  iron  and  installation  could  then  be  found, 
and  if  the  interest,  depreciation,  taxes,  and  insurance  on  this 
cost  is  just  equal  to  the  amount  saved  per  year,  there  would 
be  no  economic  advantage  in  putting  in  the  larger  pipe.  If 
the  amount  is  less  than  the  saving,  a  still  larger  pipe  should 
be  tried,  while,  if  the  saving  is  less,  it  would  be  well  to  make 
an  investigation  with  a  smaller  pipe  to  see  if  the  gain  in  interest, 
depreciation,  taxes,  and  insurance  of  a  smaller  pipe  would  not 
be  greater  than  the  increased  cost  of  power.  This  does  not 
consider  the  development  for  future  service,  which  would  alter 
the  problem. 

Having  the  area  Ad  of  the  discharge  pipe,  its  length,  and 
the  bends  in  it,  the  resistance  from  such  a  line  can  be  found, 
and  from  it  the  size  of  the  air  chamber  for  the  discharge.  Eqs. 
(86)  and  (87)  are  used  for  this. 

NUMBER  OF  REVOLUTIONS 

Ratio  of  L   to  D 

These  two  quantities  are  mutually  dependent.  In  many 
cases  the  quantity  zLN  or  piston  speed  is  the  quantity  assumed, 
and  then  D  is  known  from  the  formula 


D  = 


where  n  is  the  number  of  active  strokes  in  one  revolution. 
With  2LN=S  as  soon  as  N  is  known  L  is  given  by 

L_A 

~2N' 

The  speed  of  the  piston,  S,  is  not  a  constant,  but  may  vary 
from  50  to  700  feet  per  minute.  The  equation  shows  the 
common  tendency,  however,  to  cut  down  the  stroke  as  the 
number  of  revolutions  increases.  The  piston  speeds  used  with 
steam  engines  vary  from  about  350  to  800  feet  per  minute, 


256  PUMPING  MACHINERY 

and  where  the  pump  piston  is  mounted  in  tandem  with  the 
steam  piston,  the  piston  speed  for  the  pump  would  fix  that 
for  the  engine. 

The  length  of  the  machine  may  determine  the  stroke  to  be 
used.  Where  a  short  machine  is  desired  for  any  given  reason, 
%  whether  the  pump  be  horizontal  or  vertical,  a  short  stroke  is 
chosen. 

This  choice  of  a  short  stroke,  however,  may  mean  a  large 
diameter  and  with  it  much  heavier  parts,  cylinders,  piston 
rods,  cross-heads,  connecting  rods,  pins,  and  other  parts.  No 
general  rule  can  be  given;  each  problem  arising  must  have  its 
own  solution. 

The  value  of  N  is  determined  by  many  things.  Until 
within  the  last  twenty  years  pumps  were  usually  run  at  from 
20  to  80  revolutions  per  minute,  and  at  the  higher  speeds 
trouble  was  experienced.  In  later  years  by  increasing  the  valve 
area  or  the  number  of  valves,  and  in  cases  using  positively 
actuated  valves,  by  giving  large  passages  and  cylinder  space, 
and  by  the  use  of  large  air  chambers  much  higher  speeds  of 
revolution  were  used.  As  was  mentioned  earlier,  Riedler  was 
one  of  the  first  to  design  high  rotative  speed  pumps.  The 
speeds  have  been  carried  up  to  250  R.P.M.  By  using  such 
speeds  it  has  become  possible  to  operate  pumps  by  direct 
connection  to  electric  motors  and  gas  engines  without  the 
use  of  intermediate  gears.  Such  arrangements  save  room, 
although  on  account  of  the  special  design  for  such  pumps  the 
cost  may  not  be  reduced. 

In  America  the  practice  is  to  operate  large  pumps  with 
piston  speeds  of  about  500  feet  per  minute,  and  a  rotative  speed 
of  25  R.P.M.  This  selection  has  its  advantage,  as  the  inertia 
forces  vary  as  the  square  of  the  number  of  revolutions,  and  as 
the  first  power  of  the  crank  radius  or  stroke.  With  smaller 
pumps  60  R.P.M.  to  80  ^R.P.M.  is  often  found,  and  with  special 
pumps  the  higher  speeds  of  150  to  200  R.P.M.  are  used.  The 
latter  express  pumps  are  of  value  where  direct  connection  is 
necessary,  and  space  or  weight  limited. 

Having  then  the  size  of  the  cylinders,  revolutions,  size  of 


DYNAMICS  OF  WATER  END  257 

suction  and  discharge  pipes,  the  number  of  valves  should  be 
determined,  and  then  the  capacity  of  the  air  chamber. 

TABLE   OF  SYMBOLS 

A  =  area  of  piston  in  sq.  ft. 
ylp  =  area  of  passages  beneath  valves  at  one  end  of 

cylinder  in  sq.  ft. 

'A\  =  area  of  valve  at  one  end  of  cylinder  in  sq.  ft. 
AI,  A2,  AS,  etc.  =area  of  pipes  in  sq.  ft. 

C  =  velocity  at  air  chamber  in  ft.  per  sec.  and  con- 
stant for  spring. 
D  =  diameter  of  cylinder  in  ft. 
F  =  difference  in  discharge  from  air  chamber. 
Fs  =  pressure  from  spring  in  pounds. 
H  =  head  in  feet. 
Li,  £2  =  lengths  of  pipes  in  feet. 
L  =  length  of  stroke  in  feet. 
N  =  number  of  revolutions  per  minute. 
N'  =  number  of  effective  strokes  in  /  seconds. 
N"  =  number  of  effective  strokes  in  i  second. 
Q  =  quantity  of  water  in  cu.  ft.  discharged  in  t  sec. 
R  =  radius  of  pipe  bends  in  feet. 
5"  =  velocity  of  piston  in  feet  per  sec. 
S'  =  space  passed  over. 
V  =  volume  in  cu.  ft. 
Wv  =  weight  of  valves  in  pounds. 
a  =  acceleration  in  ft.  per  sec.  per  sec. 
b  =  equivalent  head  for  weight  and  spring  pressure. 
bi  =  breadth  of  valve  seat  in  ft. 
d  =  diameter  of  piston  rod  at  stuffing-box  in  feet 

also  diameter  of  valve  in  ft. 

di,  d2,  d%  =  diameters  of  pipe  in  ft.  (see  below  also). 
d'  =  amount  face  valve  is  closed. 
d\  —  diameter  of  passage  below  valve  in  ft. 
d2  =  pitch  of  valves  in  ft. 
e  =  excess  or  deficiency  area. 


258  PUMPING  MACHINERY 

f= friction  factor  for  flow  in  pipes. 
g  =  acceleration  of  gravity  in  feet  per  sec.  per  sec. 
h  =  head  in  ft.  of  water  and  lift  of  valve  in  ft. 
ha  =  head  in  ft.  corresponding  to  atmospheric  pres- 
sure. 

hdv  =  head  loss  at  opening  of  discharge  valve. 
hsv  =  head  loss  at  opening  of  suction  valve. 
hs  =  pressure  head  on  suction  side  of  pump  piston. 
hd  =  pressure  head  on  discharge  side  of  pump  piston. 
fi2  =  height  in  ft.  from  fore  bay  to  lower  end  of  pis- 
ton stroke. 
//2'  =  similar  height  on  discharge  side  from  end  of 

piston  stroke, 
fe  =  head  lost  in  friction,  velocity,  enlargement,  etc., 

in  ft. 

h±  =  head  due  to  inertia  in  ft. 
h$  =  head  from  level  in  suction  air  chamber  to  end 

of  stroke. 
hs  =  head  from  level  in  discharge  air  chamber  to  end 

of  stroke. 

hp  =  hea.d  loss  in  pipe  measured  in  feet  of  water. 
i  =  number  of  ribs  in  valve  seat. 
k  =  constant  for  pipe,  pipe  bends. 
/  =  length  of  pipe  in  ft.  and  periphery  of  valve. 
m  =  constant  for  valve  friction. 
n  =  ratio  of  length   of   connecting   rod    to  crank 

radius. 

wc  =  number  of  effective  strokes  in  i  rev. 
n'  =  number  of  valves  at  one  end. 
pc  =  pressure  below  valve  in  Ibs.  per  sq.  ft. 
pu  =  pressure  above  valve  in  Ibs.  per  sq.  ft. 
q  =  discharge  in  cu.  ft.  per  sec. 
r  =  crank  radius  in  ft. 
s  =  thickness  of  ribs  at  end  or  lap  of  multiported 

valve. 

s'  =  space  passed  over. 
t  =  time  in  seconds. 


DYNAMICS  OF  WATER  END  259 

z>  =  velocity  of  water  in  ft.  per  sec.,  velocity  under 

valve  edge. 

•D1  =  velocity  of  water  in  valve  passage. 
D2  =  velocity  of  valve. 
vs  =  velocity  in  smaller  pipe  where  sections  change 

suddenly. 

w  =  weight  of  i  cu.  ft.  of  water  in  Ibs. 
x  =  movement  of  piston  or  cross  head  from  head 

end  of  stroke  in  feet  and  pitch  of  rings  in 

multiported  valve. 
^2  =  lifts  in  ft. 
yo  =  compression  of  spring. 
a  =  angular  inclination  of  connecting  rod. 
a  =  coefficient  in  Bach's  formula. 
|8  =  coefficient  in  Bach's  formula. 
7  =  coefficient  in  Bach's  formula. 
7  =  coefficient  of  discharge. 

6  =  angular  motion  of  crank  from  head  dead  center. 
0  =  inclination  of  cylinder. 
p  =  ratio  of  diameter  of  small  pipe  to  that  of  large 

pipe. 

ju  =  coefficient  for  valve  velocity. 
£  =  coefficient  for  lost  head,  coefficient  for  valve 

loss,  coefficient  for  valve  discharge, 
co  =  angular  velocity.     •  { 


CHAPTER  VI 
DESIGN  OF  PARTS 

WATER  CYLINDERS 

AFTER  the  diameter  and  stroke  have  been  determined,  the 
cylinder  is  designed.  The  arrangements  -of  the  bore,  the 
passages  leading  to  the  valves,  the  valves,  the  valve  decks, 
and  the  valve  chests  are  all  of  importance,  and  are  varied 
according  to  the  peculiarities  of  the  designer.,  The  general 
principle  is  to  make  the  path  of  the  water  as  direct  as  possible; 
the  construction  of  the  casting  simple,  and  the  position  of  the 
valves  such  that  they  may  be  easily  examined  and  replaced. 
To  show  the  arrangements  proposed  and  used  by  pump  designers 
a  number  of  typical  pump  cylinders  have  been  chosen,  each 
illustrating  some  special  point.  Although  a  few  might  have 
answered  the  purpose,  the  number  has  been  increased  to  famil- 
iarize the  student  with  the  construction  of  modern  pumps, 
to  show  many  ways  of  constructing  machines,  and  to  illustrate 
how  constructions  may  be  simplified. 

The  water  pistons  take  one  of  four  principal  forms:  ist, 
the  piston  (A,  Fig.  199);  2d,  the  plunger  and  ring  (B,  Fig.  199); 
3d,  the  packed  plunger  (C,  Fig.  199),  and  4th,  the  bucket 
(D,  Fig.  199). 

These  are  packed  in  different  ways.  Fig.  199^  .  shows  a 
canvas  packing  inserted  on  a  ledge  formed  on  the  piston  casting. 
The  packing  is  usually  made  of  layers  of  canvas  or  square- 
woven  cotton  indurated  with  rubber,  or  it  may  be  a  square 
flax  packing.  The  form  of  packing  shown  at  B  has  been 
used  for  many  years;  the  long  sleeve  makes  practically  a 
water-tight  joint,  and  for  clear  water  this  lasts  for  some  time. 
The  sleeve  may  be  renewed  when  necessary.  The  plunger 
(C,  Fig.  199)  is  packed  on  the  inside  at  the  center.  Such  packing 

260 


DESIGN   OF   PARTS 


261 


may  be  placed  at  the  outside.  The  form  is  identically  the  same 
in  both  cases.  The  kind  of  packing  used  is  the  same  as  that 
employed  on  the  piston.  The  cup-leather  packing  (D,  Fig.  199) 
is  used  with  deep-well  pumps  and  with  high-pressure  pumps. 
The  cup  leather  forms  a  good  packing  in  such  cases,  as  the 


B  D 

FTG.  199. — Pistons,  Plungers  and  Bucket. 

pressure  exerted  by  the  leather  varies  with  the  water  pressure, 
and  so  the  friction  on  the  suction  stroke  is  small.  Fig.  200 
shows  the  form-  for  a  double-acting  piston  with  cup-leather 
packing.  When  packings  are  used  for  the  outside  they  are 
practically  the  same  in  form.  The  piston  rod  is  packed  as 
shown  in  Fig.  201  or  Fig.  202;  the  plunger  for  ordinary  pres- 


262 


PUMPING  MACHINERY 


sures  with  a  light  packing  (Fig.  203),  while  for  heavy  pressures  in 

pumps  or  hydraulic  jacks  U-leather  or 
hydraulic  packing  (Fig.  204)  is  used. 

The  design  of  these  packings  is 
usually  empirical,  and  the  figures 
shown  are  marked  with  numbers 
which  are  given  in  terms  of  a  unit. 

In  Figs.  201  and  202  the  propor- 
tional unit  is  \d,+y  where  d=  diam- 
eter of  piston  rod. 

In  Fig.  203  the  size  and  number 
of  the  bolts  and  the  thickness  of  the 
stuffing  box  will  have  to  be  designed. 
The  thickness  A  of  the  packing  and 
the  depth  B  depend  to  a  certain  ex- 
tent on  the  size  of  the  plunger.  B=o.2  to  0.4^  and  A  =  o.id 


FIG.  200. — Double-leather 
Packing. 


"  up  to  i"  may  be  used  as  a  guide. 


FIG.  20 1. — Stuffing  Box. 

The  thickness  C  is  designed  for  high  pressures  as  a  cylinder 
wall  under  the  pressure  of  the  fluid. 

The  bolt  area,  a,  or  the  number  of  bolts,  n,  may  be  found 
by  the  following  formula  if  one  of  these  quantities  is  assumed: 


DESIGN  OF  PARTS 


263 


The  cup  leathers  (Fig.  206)  are  usually  made  by  soaking 
the  leather  to  make  it  pliable,  and  then  forcing  it  into  the 
mold  (Fig.  206)  by  a  press  or  bolt.  After  it  is  forced  down 


FIG.  202. — Stuffing  Box. 

it  is  allowed  to  dry  and  set,  when  it  is  trimmed  on  the  edge  by 
turning.  The  wear  on  cup  and  other  leather  packing  occurs 
near  the  bend  in  the  leather  (Fig.  205)  as  it  is  this  point  which 


FIG.  203. — Plunger  Packing. 


FIG.  204. — U-Leather  Packing. 


is  driven  against  the  plunger  or  cylinder  wall.  Since  this  is 
the  case  there  is  no  necessity  for  making  the  dimensions  h 
greater  than  the  amount  given  below. 

The  U-leathers  are  proportioned,  as  shown  in  Fig.  207,  for 


264 


PUMPING   MACHINERY 


FIG.  205. — Wear  in  Cup  Leather. 


FIG.  207. — U  Leather. 


FIG.  206. — Formation  of  Cup  Leather. 


FIG.  209. — Hat-leather  Packing. 


FIG.  208. — Formation  of  U  Leather.      FIG.  210.— Formation  of  Hat  Leather. 


DESIGN  OF  PARTS  265 

which  the  various  dimensions  are  given  later.  These  are 
made  by  formers  or  molds,  after  soaking,  as  shown  in  Fig. 
208.  This  operation  may  best  be  performed  on  a  hydraulic 
press.  The  first  fold  is  made  as  in  the  second  part  of  the  figure, 
after  which  the  last  core  is  introduced. 

The  hat  leather  used  as  shown  in  Fig.  209  is  formed  in  the 
same  manner  as  the  cup  and  U  leathers  by  formers  (Fig.  210). 
After  drying,  the  center  is  cut  out  and  the  edge  chamfered. 

For  cup  leathers  (Fig.  205), 

A  =  i"to  i"; 


Friction  =cdp\ 

0=0.03  to  0.05; 

p=  hydrostatic  pressure  per  sq.  in. 

For  U  leathers  (Figs.  204,  and  207), 

Diameter  box  =  diameter  plunger  +  2w  —  J"  ; 

Outside  diameter  leather  =  diamet  er  plunger  +  2  w—^" 
Inside  diameter  leather     =  diameter  plunger 
Height  h  =  1.6  width  =£"  to 

Width  =w=yto%"; 

-\(D 

Bolt  area  =a=—  -^ 

n^t 

Height  of  flange  A"  =  1.5  diameter  bolts; 

h'  =  3  diameter  bolts  ; 
Width  of  cylinder  flange   =3  diameter  bolts; 

P 

Friction  =0.04-3  =  o 


The  Marsh  pump  (Fig.  211)  has  one  of  the  simplest  forms 
of  water  end.  In  this  design  there  is  no  lining  to  the  cylinder 
bore.  The  suction  chamber  A  is  formed  by  coring,  and  the 
discharge  chamber  B  is  formed  by  the  valve  deck  plate  C  and 
the  cover  D.  The  valve  seats  are  forced  into  holes  on  the 
valve  decks.  The  piston  is  packed  with  cup  leathers.  This 


266 


PUMPING  MACHINERY 


particular  pump  is  built  for  pumping  milk  and  for  that  reason 
it  is  so  constructed  that  it  may  be  taken  apart  quickly  and 
easily  for  cleaning.  All  nuts  are  wing  nuts,  and  the  bolts  are 
hinged  so  that  they  may  be  swung  out.  The  piston  rod  is 
light  and  the  stuffing  box  is  simple. 

Another  form  of  Marsh  pump  (Fig.  212)  is  so  constructed 
that  the  cylinder  bore  is  fitted  with  a  brass  liner  to  take  the 
wear  of  the  piston.  The  liner  is  held  in  place  by  screws  passing 
through  a  flange.  It  may  be  renewed  easily  when  necessary. 
The  piston  body  F  is  fastened  to  the  piston  rod  by  a  nut,  while 


FIG.  211. — Marsh  Milk  Pump. 

the  follower  plate  G  is  held  on  the  same  thread  by  another  nut. 
The  removal  of  this  plate  permits  one  to  examine  or  repair 
the  piston  packing.  The  shoulder  on  the  piston  rod  holds 
the  body  F  in  place,  preventing  any  play. 

The  stuffing  box  H  is  of  the  cap  form,  and  is  made  separate 
from  the  cylinder  casting.  This  simplifies  the  foundry  work 
and  lessens  the  difficulty  of  the  machine  work  in  the  manu- 
facture. 

The  method  of  dividing  the  valve  chambers  of  the  cylinder 
ends  by  a  partition  carried  to  the  discharge  valve  deck  as  well 
as  the  method  of  attaching  the  air  chamber,  and  the  form  of 


DESIGN  OF   PARTS 


267 


the  air  chamber  are  all  to  be  noted.  The  cap  on  the  end  of 
the  air  chamber  has  been  employed  so  that  a  core  print  may 
be  used  at  each  end  of  the  pattern. 


FIG.  212. — Marsh  Pump. 

To  simplify  the  water  end,  the  Marsh  Company  build  a 
pump  (Fig.  213)  in  which  the  water-cylinder  bore  is  made  by 


268 


PUMPING  MACHINERY 


using  a  piece  of  solid-drawn  brass  pipe.  This  makes  a  very 
simple  casting,  as  the  suction  chamber  A  is  carried  up  from 
the  base  of  the  pump.  The  cap  forming  the  discharge  chamber 


•  FIG.  213. — Marsh  Pump. 

B  is  cast  solid  with  the  air  chamber.  The  valves  used  in  this 
pump  are  peculiar  in  form  and  the  guides  are  so  arranged  that 
the  valves  are  not  subject  to  a  lifting  pressure  after  they  rise 
to  a  certain  height.  The  valve  seats  are  inserted  in  openings 


DESIGN  OF  PARTS 


269 


^tfttf^^ 

FIG.  214. — Marsh  Pump. 


in  the  valve  decks,  simplifying  the  casting  and  making  repairing 
easy.     The  peculiar  form  of  packed  piston,  the  stuffing  box, 


270  PUMPING  MACHINERY 

and  the  simple  form  of  casting  are  to  be  observed  in  the 
figure. 

For  larger  sizes  the  Marsh  pump  water  end  is  built  as  shown 
in  Fig.  214.  In  this  case  the  suction  chamber  is  a  separate 
casting,  and  forms  the  base  of  the  water  end.  The  suction 
valve  deck  is  reinforced  against  the  water  pressure  from  the 
cylinder  by  a  rib  5.  Such  a  construction  is  necessary,  as  the 
valve  deck  is  a  broad,  flat  plate,  subject  to  a  downward  pres- 
sure. A  large  suction  pipe  enters  the  side  of  the  suction 
-chamber.  The  discharge  valve  deck  is  part  of  the  cylinder 
proper,  and  is  stiffened  by  the  rib  T  which  extends  across  the 
casting,  while  the  discharge  chamber  is  bolted  to  the  cylinder. 
The  air  chamber  is  attached  to  this  and  is  similar  in  form  to 
those  used  on  the  smaller  pumps.  Hand  holes  are  arranged 
in  the  side  of  the  cylinder  casting,  and  in  the  discharge  chamber 
at  MM,  so  that  the  valves  may  be  inspected  or  repaired  without 
the  necessity  of  removing  the  large  castings  or  opening  large 
covers.  Such  construction  is  always  necessary  in  pumps  of 
any  size. 

The  valves  are  of  a  different  form  from  those  shown  before. 
These  are  spring-controlled  valves.  Their  action  is  similar 
to  that  of  the  others  illustrated.  The  seats  are  made  separate, 
a  good  construction,  in  order  that  seats  may  be  renewed  if 
broken,  and,  moreover,  the  casting  is  thereby  simplified.  The 
follower  ring  R  on  the  piston  replaces  the  plate  used  on  the 
other  pistons,  and  the  bolted  gland  in  the  stuffing  box  replaces 
the  cap  gland.  The  liner  of  brass  is  used  to  simplify 
renewal  when  ,the  wear  from  the  piston  packing  becomes 
excessive. 

One  of  the  important  features  in  pump  design  is  to  have  an 
ample  and  simple  direct  passage  for  the  water  through  the 
pump.  The  pumps  shown  in  the  preceding  figures  have  been 
good  in  this  respect,  but  the  Fairbanks-Morse  pump  in  Fig. 
215  shows  in  a  somewhat  better  manner,  the  ample  passages. 
The  suction  enters  the  chamber  A  in  an  easy  sweep  from  D, 
and  then  goes  through  a  large  passage  to  the  pump  cylinder. 
The  discharge  from  the  other  side  is  forced  into  B  and  leaves 


DESIGN  OF  PARTS 


271 


through  a  passage  to  the  discharge  E.    The  general  features 
of  this  design  are  seen  on  inspection  of  Fig.  215. 

Another  pattern  with  large  passages  is  shown  in  the  Worth- 
ington  pump  (Fig.  216).  In  both  of  these  pumps  the  cylinder 
bore  is  lined  by  using  a  piece  of  brass  tubing.  This  tubing  is 
withdrawn  when  worn,  and  another  piece  inserted  in  its  place. 
In  pumps  intended  to  lift  acids  which  may  attack  the  metal, 


FIG.  215. — Fairbanks-Morse  Pump. 

. 

a  lining  of  wood  has  been  used  for  the  bore  of  the  cylinder, 
and  the  valve  chambers  as  well  as  the  piping. 

.  In  Fig.  216  the  peculiar  form  of  follower  ring  or  plate  is  to 
be  noted.     This  gives  a  very  long  wearing  surface. 

The  Worthington  plunger  and  ring  form  of  water  end  (Fig. 
217)  is  one  which  has  many  advantages.  The  water  enters 
the  suction  chamber  A  through  C  and  passes  up  into  the  pump 
cylinders,  and  from  there  it  is  forced  into  the  discharge  pipe 


272 


PUMPING  MACHINERY 


D  through  the  chamber  B.  The  plunger  is  fastened  to  the 
piston  rod  by  two  nuts.  It  operates  through  a  sleeve  E,  which 
is  held  in  the  partition  between  the  sides  by  a  ring  F.  The 
sleeve  is  called  a  ring.  There  is  no  tight  packing.  The  joint 


FIG.  216. — Worthington  Pump. 

is  so  long  that  there  is  practically  no  leakage.     This  type  of 
pump  has  been  quite  successful. 

Outside-packed  plunger  pumps  have  the  advantage  that 
the  leakage  past  their  displacing  parts  is  visible.  They  are  of 
various  forms.  The  piston  rod  is  at  times  carried  through  the 
two  plungers  (Fig.  218),  the  rod  passing  through  a  sleeve  or 
stuffing  box.  In  this  pump  the  suction  enters  at  A  and  is 


DESIGN  OF   PARTS 


273 


carried  to  the  valves,  and  finally  discharges  into  the  chamber 
C  and  from  it  into  B.     This  type  of  end  outside-packed  plunger 


FIG.  217. — Plunger  and  Ring 


FIG.  218. — Worthington  Pump. 

pump  is  simple,  and  the  water  passages  are  all  ample.  The 
great  disadvantage  is  in  the  possibility  of  leakage  around  the 
piston  rod.  To  obviate  this  leakage  the  plungers  are  -con- 
nected on  the  outside  by  a  trombone  frame  (Fig.  219).  The 


274 


PUMPING  MACHINERY 


rods  E  and  F  join  the  two  cross  heads  G  and  H  together.  This 
pump  may  be  the  same  in  form  as  Fig.  218,  with  the  exception 
of  the  connecting  rod  and  sleeve. 

A  center  outside-packed  pump  (Fig.  220)  does  away  with 


FIG.  219. — Outside-packed  Plungers. 


FIG.  220. — Center  Outside  Packing. 

the  necessity  of  outside  rods  and  gives  a  pump  in  which  all 
leakage  past  the  plungers  or  rods  is  visible;  the  stuffing  boxes 
are  also  clearly  shown. 

Water  enters  the  suction  chamber  A  from  B,  passes  directly 
into  the  pump,  and  is  driven  into  chamber  C,  and  from  there 


DESIGN  OF  PARTS 


275 


into  the  discharge  D.  The  space  F  is  opened  to  the  outside  so 
that  the  stuffing  boxes  are  always  visible.  As  constructed  in 
this  figure  the  chambers  C  and  A  are  cast  solid  with  the  other 
parts.  This  is  riot  always  the  method  of  construction,  as  in 
many  cases  the  two  parts  of  the  water  end  are  separate  castings 
connected  by  the  chambers  A  and  C. 

The  discharge  valves  are  placed  beneath  hand  plates  G 
and  H,  while  the  suction  valves  may  be  examined  and  repaired 
through  side  hand  holes  in  the  pump  barrel.  The  valves  have 
removable  seats  and  these  are  so  placed  that  there  is  a  direct 


SECTION  THROUGH  x-y 


FIG.  221. — Fire  Pump. 


path  of  sufficient  dimensions  through  the  pump.  Such  arrange- 
ments  cut  down  the  friction  and  give  a  more  smoothly  acting 
pump. 

These  same  points  of  a  direct  path,  simplicity  and  ease  of 
repair,  are  seen  in  the  double-acting  fire  pump  (Fig.  221), 
where  the  water  enters  at  A  and  leaves  at  B.  The  piston  is 
packed  with  cup  leathers,  and  the  passages  from  it  to  the 
suction  valves  are  very  large.  The  reason  for  this  is  the  high 
rotative  speed  of  these  pumps  and  the  necessity  for  having  no 
interruption  in  the  water  column.  There  is  not  the  need  for 
such  large  passages  on  the  discharge  side  as  the  water  is  being 
forced  out.  By  removal  of  certain  parts,  these  valves  may  be 


276 


PUMPING  MACHINERY 

A 


FIG.  222. — Metropolitan  Fire  Pump. 


DESIGN  OP  PARTS 


277 


examined,  although  this  is  not  so  easily  done  as  in  most  other 
pumps.  These  pumps  are  so  closely  built  on  account  of  the 
lack  of  room  that  many  desirable. features  have  to  be  sacrificed 
in  order  that  more  urgent  needs  may  be  met.  The  Ahrens 


FIG.  223. — Railroad  Pump. 

pump  has  been  improved  by  the  American  Fire  Engine  Co. 
in  their  Metropolitan  engine.  The  suction  valves  (Fig.  222) 
have  been  replaced  at  each  end  of  the  cylinder  through  five 
suction  valves  of  large  area.  The  two  discharge  valves  have 
not  the  same  area  as  the  suction,  but -friction  here,  although 
costly,  will  not  interfere  with  the  smooth  running  of  the  pump. 


278  PUMPING  MACHINERY 

The  discharge  occurs  through  B.  C  is  the  discharge  air  chamber 
and  D  is  the  suction  air  chamber. 

The  valves  of  the  pump  are  easily  examined  or  replaced 
by  the  removal  of  the  cylinder  heads  EE  or  hand-hole  plate 
-F  over  the  discharge  valves.  When  necessary  the  cylinder 
liner  may  be  renewed.  The  piston  is  made  deep  and  con- 
tains a  number  of  grooves. 

Fig.  223  is  a  detail  of  the  water  end  of  a  pump  used  for 
railroad  work.  It  is  driven  by  a  gas  engine  as  shown  in  Fig. 
147.  There  are  four  valves  arranged  around  the  pump  barrel; 
two  suction  and  two  discharge.  The  suction  valve  shown  in 
the  figure  is  connected  with  the  upper  end  of  the  cylinder, 
while  the  discharge  valve  is  connected  with  the  lower  end. 
The  valves  not  shown  are  connected  with  the  opposite  ends. 
The  piston  rod  A  is  guided  by  the  arm  B  while  the  cross-head 
C  is  connected  with  the  pitman  bar  from  a  pin  on  a  gear  wheel. 
The  air  chamber  D  serves  to  steady  the  discharge. 

The  valve  boxes  are  provided  with  covers  which  are  held 
in  place  by  yoke  pieces  so  that  the  valves  may  be  easily  exam- 
ined. The  valve  boxes  are  of  ample  size. 

The  barrels  of  deep-well  pumps  (Fig.  224)  are  usually  made 
of  a  piece  of  brass  tubing  lowered  into  the  well.  The  valves 
are  of  the  disc-lift  type  or  the  ball  type.  The  foot  or  suction 
valve  is  either  lowered  into  place  and  held  to  its  seat  by  its 
own  weight  or  it  may  be  put  in  place  by  attaching  it  to  the 
pump  rods  and  forcing  it  into  position.  It  is  necessary  to 
have  the  foot  valve  water  tight,  as  there  will  be  lifting  should 
the  water  be  forced  around  it.  These  foot  valves  and  buckets 
are  usually  packed  with  cup  leathers. 

The  pump  rods  are  often  made  of  close-grain  lumber  with 
iron  armored  ends.  One  end  is  made  into  a  nut,  the  other 
end  is  threaded  as  a  bolt.  In  threading  these  ends  considerable 
taper  is  used,  so  that  it  is  only  necessary  to  turn  the  rod  three 
or  four  times  to  have  ten  or  fifteen  threads  in  contact  when 
joining  the  sections.  Precautions  must  be  taken  to  prevent 
these  screws  from  backing  off,  as  the  pump  line  depends  on 
their  holding. 


DESIGN  OF  PARTS  279 

Pressure  pumps  (Fig.  225 )  are  made  with  heavy  .cylinder 


FIG.  224. — Deep-Well  Pump  and  Rods. 


y 

A 


y 


walls,  and  in  most  cases  they  are  of  the  plunger  type,  as  this 
design  is  very  suitable  to  high  pressures.     The  packing  may  be 


280 


PUMPING   MACHINERY 


DESIGN    OF  PARTS  281 

of  the  ordinary  hemp  form  shown  in  the  figure  or  leather  packing 
may  be  used.  The  valve  chambers  are  usually  single  castings. 
They  vary  in  form;  sometimes  the  casting  contains  both  suc- 
tion and  discharge  valves,  and  at  other  times  only  one  valve 
is  in  each  casting.  The  valve  castings  A  A  are  bolted  to  the 
suction  casting  B,  and  the  discharge  casting  C  as  well  as  to  the 
pump  barrels  DD.  This  arrangement  gives  a  simple  pump 
cylinder,  and  although  the  valve  chambers  are  not  simple, 
they  are  easily  built,  and  machined.  The  valves  are  so  placed 
that  they  may  be  examined  and  repaired  by  the  removal  of  a  cap. 
When  these  pressure  pumps  are  made  larger  and  more  valves  are 
required,  the  valve-box  castings  are  increased  in  number. 


FIG.  226. — Pressure  Pump. 

In  smaller  pumps  under  very  high  pressure  the  valve 
chambers  may  be  cast  with  the  pump  barrel  as  shown  in  Fig. 
226,  which  represents  a  high-pressure  Burnham  pump.  The 
small  valves  are  so  arranged  that  by  removing  one  plug  both 
suction  and  discharge  valves  may  be '  examined.  Attention 
is  called  to  the  great  difference  between  the  plunger  and  the 
steam  piston  areas  and  to  the  thickness  of  the  cylinder  wall. 
The  valve  seats  are  renewable,  so  that  if  any  of  these  should 
wear  it  could  easily  be  replaced  without  disconnecting  the 
pump. 

The  water  end  of  large  pumps  has  been  greatly  simplified. 
One  of  the  older  pumps  for  the  Boston  sewage  system,  designed 
by  Mr.  E.  D.  Leavitt,  Jr.,  is  shown  in  Fig.  227.  In  this  case 
the  valves  were  rectangular  clack  valves,  3f  Xi3i  inches.  Six 


282 


PUMPING  MACHINERY 


of  them  were  attached  to  a  frame  and  placed  on  the  openings  of 
the  suction,  while  three  discharge  valves  were  attached  to  ea.ch 
frame  on  the  other  side.  The  frame  with  the  suction  valves 

open  as  would  occur  on  the 
suction  stroke  is  shown  on 
the  left,  while  on  the  right 
the  frame  alone  is  to  be 
seen  in  section  and  at  the 
upper  right  hand  the  dis- 
charge valves  are  shown 
closed. 

The  suction  valves  and 
discharge  valves  are  in  the 
proportion  of  36  to  27.  The 
reason  for  this  was  seen  in 
Chapter  V.  In  all  cases  an 
endeavor  is  made  to  cut 
down  the  losses  on  the 
suction  side.  The  clear 
passage  to  and  from  the 
valves  is  necessary,  and 
valves  of  large  areas  with 
no  supports  obstructing  the 
passage  are  required  to 
pass  the  solid  matter  which 
is  found  in  sewage.  Mr. 
Leavitt  states  that  there 
is  a  record  of  this  pump 
having  passed  through  its 
water  cylinder  a  plank 
2X12X36  inches. 

The  water  end  is  48 
inches  in  diameter  and  of 

form 


FIG.  227.-Sewage  Pump. 

of  the  plunger  is  dotted  in  the  lowest  position.  The  form  of 
stuffing  box  is  illustrated  together  with  the  grooves  on  the 
inside  of  the  gland  and  bushing.  This  form  of  so-called 


DESIGN  OF  PARTS 


283 


labyrinth    packing    has    been    proven    of    little    value.     The 
manhole   in   the   center  ring  of    the  pump   permits  entrance 


FIG.  228. — Lawrence  Pump. 

for  the  examination  of  the  suction  valves,  while  a  manhole 
in  the  discharge  chamber  permits  the  examination  of  the  other 
valves.  The  cylindrical  casting  is  made  in  three  principal 


284 


PUMPING  MACHINERY 


parts  to  simplify  the  foundry  work  and  to  make  shipping  arid 
erection  less  difficult. 


FIG.  229. — Ontario  Pump. 

Fig.  228  illustrates  the  Lawrence  Water  Works  pump  of 
Mr.  E.  D.  Leavitt,  Jr.     It  is  the  bucket  form  of  pump  with  a 


DESIGN  OF  PARTS 


285 


supplementary  discharge  at  A  for  the  purpose  of  giving  a 
passage  to  the  water  should  the  valve  in  the  bucket  cease  to 
operate.  When  the  pump  operates  as  a  bucket  pump  the 
course  of  the  water  is  direct  through  the  pump  barrel.  The 
use  of  a  plunger  enlargement  on  the  rod  of  one-half  the  area 
of  the  piston  serves  to  produce  a  discharge  on  each  stroke, 
while  suction  occurs  on  every  other  stroke. 

The   valves   for  this   pump   were   originally   1 6-inch  brass 
double-beat  valves  shown  in  Fig.  252,  but  as  the  friction  of  the 


/; 

s5< 

^ 

1 

FIG.  230. — Milwaukee  Pump. 

pump  was  excessive  the  valves  were  changed  to  the  annular 
form  as  shown  in  Figs.  228  and  250.  This  reduced  the  friction 
very  materially.  The  castings  making  up  the  pump,  the  air 
chamber  B,  the  manholes,  valve  decks  and  other  details  are 
seen  by  a  study  of  the  picture. 

A  single-suction  double-discharge  plunger  pump  is  shown 
in  Fig.  229.  In  this  type  of  pump  for  Ontario,  Leavitt  used  a 
large  number  of  small  valves  and  on  the  suction  side  small 
draft  tubes  were  placed  below  the  valves.  The  plunger  was 


286 


PUMPING   MACHINERY 


shaped  to  fit  into  the  bottom  of  the  cylinder  and  cause  all  of 
the  water  to  be  in  circulation.  Such  a  construction  may  be 
questioned  unless  there  is  a  lack  of  head  room  and  it  is  thought 
necessary  to  point  the  plunger.  The  plunger  has  an  extension 
sleeve  on  it  of  one-half  the  area  of  the  plunger.  The  castings 
are  simple  and  well  braced  by  brackets  and  webs,  as  well  as 


FIG.  231. — Cincinnati  Pump. 

reinforced  by  rings  around  the  barrel.  The  path  of  the  water 
is  direct,  and  the  head  pressure  on  the  discharge  valve  deck  is 
supported  by  the  webs  A  A,  which  act  as  girders. 

The  valves  are  spring  controlled,  of  small  size,  and  shown  in 
greater  detail  in  Fig.  252. 

The  Allis  pump  of  Milwaukee  (Fig.  230)  is  .another  case  in 
which  a  series  of  lift  valves  were  used  to  replace  large  double- 
beat  valves.  In  this  case,  each  large  valve  was  replaced  by  a 
box  or  cage  on  the  sides  and  top  of  which  were  openings  for 


DESIGN  OF  PARTS 


287 


small  valves  as  shown  in  detail  in  Fig.  254.  The  construction 
shown  in  Fig.  230  illustrates  how  simple  the  castings  of  a  large 
purnp  may  be.  The  valve  decks,  which  are  reinforced  by  cross 
ribs  or  girders  to  withstand  the  pressure,  are  the  end  parts  of 
individual  castings;  on  account  of  this  the  faces  may  be  easily 
machined  in  the  shop,  and  erected  in  the  field.  The  path  of 


FIG.  232. — Snow  Pump  End. 

the  water  is  more  or  less  direct  and  the  suction  pipe  is  of  proper 
size  for  this  slow-running  pump. 

Fig.  231  illustrates  a  3o,ooo,ooo-gallon  water  end  built  by 
the  Holly  Pump  Co.  The  suction  pipe  enters  the  valve  chamber 
on  each  side  of  the  pump  at  AA\  the  valve  decks  BB  are  made 
of  similar  castings  of  considerable  depth  to  withstand  the  high 
working  pressure.  The  valves  are  mounted  on  cages  as  in  the 
previous  figure.  The  discharge  occurs  at  CC.  The  space  D 
in  the  top  of  the  discharge  acts  as  an  air  chamber.  The  pipes 
EE  are  used  to  introduce  compressed  air  into  this  space.  The 


288 


PUMPING  MACHINERY 


plunger  F  draws  water  from  the  suction  valves  on  each  side 
of  it  on  its  up  stroke  and  discharges  through  the  two  sets  of 
discharge  valves  on  its  down  stroke.  The  connection  H  made 
between  the  discharge  side  and  pump  space  is  used  to  prime 
the  suction  valves  when  necessary,  but  may  also  be  used  in 
starting  the  pump  to  equalize  the  pressure  on  the  plunger 
on  each  stroke.  A  large  suction  air  chamber  is  shown  at  K. 
Fig.  232  shows  the  water  end  of  a  large  horizontal  pump. 


FIG.  234. — Metal  Hinge  Clack  Valve. 


FIG.  233.— Leather  Clack  Valve. 


FIG.  235.— Clack  Valve. 


The  water  end  is  made  up  of  four  principal  castings:  A  suction 
chamber  A\  two  pump  ends  BB;  and  a  discharge  chamber 
C,  containing  the  air  chambers  DD.  The  valve  decks  in  the 
pump  ends  BB  contain  a  large  number  of  small  valves;  they 
are  well  braced  by  cross  ribs  as  seen  in  the  figure.  The  man- 
hole cover  at  E,  which  is  hung  from  an  arm,  serves  for  the  exam- 
ination of  the  suction  valves,  while  that  at  F  is  used  for  the 
discharge.  The  figure  illustrates  very  clearly  the  method  of 
packing  the  plunger  and  piston  rods  and  the  simple  manner  in 


DESIGN   OF  PARTS 


289 


which  the  castings  can  be  made.  The  trough  G  beneath  the 
intermediate  stuffing  boxes  is  intended  to  catch  the  drip. 

One  of  the  simplest  valves  found  in  small  hand  pumps  is 
the  leather  clack  valve  (Fig.  233).  A  circular  piece  of  leather 
shown  in  plan  in  the  figure  has  a  groove  cut  out  of  it  and  after 
fastening  two  iron  washers  to  this,  the  leather  is  held  beneath 
the  pump  barrel  and  the  valve  seat  casting.  The  small  part 
of  the  leather  left  after  cutting  the  groove  forms  a  hinge.  The 
upper  washer  is  not  only  used  to  weight  the  clack,  but  it  also 
supports  the  weight  of  water  above  the  leather,  making  a  water- 
tight valve. 

In  Fig.  234  the  leather  hinge  has  been  replaced  by  a  pin, 
and  in  Fig.  235  the  leather  hinge  is  retained  for  a  valve  which  is 


FIG.  236. — Butterfly  Valve. 

rectangular.  Such  a  valve  is  used  for  sewage  pumping  (Fig. 
227).  A  double  clack  valve  (Fig.  236)  is  sometimes  called  a 
butterfly  valve.  In  this  form  the  valve  is  made  rectangular. 
Such  a  shape  is  often  found  when  clack  valves  are  used.  The 
valve  seat  is  formed  separate  from  the  pump  casting  and  is 
bolted  in  position.  Stops  are  usually  employed  to  keep  the 
valve  from  opening  too  much. 

A  metal  clack  valve  (Fig.  237)  may  be  used  at  times  where 
the  liquid  handled  would  destroy  the  leather..  The  pivot  in 
this  case  moves  in  a  slot  to  allow  for  the  wear  of  the  valve. 
A  pin  through  a  hole  would  not  allow  the  valve  to  seat  properly 
after  wear  had  occurred. 

The  conical  valve  of  Fig.  238  is  guided  by  wings  on  its  lower 
face.  The  seat  of  this  valve  is  often  made  of  bronze,  and 


290 


PUMPING  MACHINERY 


inserted  in  the  valve  deck.  This  renders  it  an  easy  matter 
to  renew  the  seat  and  permit  one  to  use  a  better  metal  for  it, 
the  valve  deck  being  made  of  a  soft  iron.  The  valve  is  fitted 
to  the  seat  by  grinding.  The  operation  consists  in  turning  the 
valve  against  the  seat  by  a  screwdriver,  after  oil  and  emery 
are  introduced  between  them.  To  give  the  valve  in  seating 


FIG.  237.— Metal  Clack  Valve. 

a  turning  motion,  the  wings  below  the  valve  disc  are  bent  into 
a  helical  form  (Fig.  239).  The  action  of  the  water  on  the 
vanes  is  to  rotate  the  disc.  This  allows  the  valve  to  seat  at 
different  points  each  time,  thus  eliminating  the  excessive 
grooving  which  occurs  when  a  valve  always  seats  at  the  same 
point  after  the  formation  of  an  incipient  groove.  The  rotary 
motion  also  gives  the  valve  a  wiping  action  in  seating. 


n 


^^^ 

s 

/ 

\\\\\\^ 

n 


FIG.  238. — Conical  Valve. 


FIG.  239. — Helical  Wings. 


The  ball  valve  (Fig.  240)  is  an  effective  type.  The  action 
of  the  cage  over  the  ball  and  the  removable  nature  of  all  parts 
are  evident  from  the  figure. 

One  of  the  common  forms  of  valves  used  with  all  forms  of 
pumps  is  shown  in  Fig.  241.  The  valve  seat  is  made  of  com- 
position brass,  and  screws  into  the  valve  deck.  The  valve  is 
made  of  a  composition  of  rubber  and  other  substances.  It  is 
backed  by  a  piece  of  sheet  brass  against  which  the  spring  presses. 


DESIGN   OF  PARTS 


201 


The  spring  is  held  beneath  a  nut  on  the  valve  spindle.     The 
spindle  is  screwed  down  tight  against  a  shoulder  so  that  it 


FIG.  240. — Ball  Valve. 

will  not  back  off.     A  split  pin  put  in  the  hole  in  the  top  of  the 
spindle  prevents  the  nut  from  unscrewing.     All  parts  shown 


FIG.  241. — Disc  Valve. 

in  the  figure  with  the  exception  of  the  disc  are  made  of  brass. 
The  figure  shows  the  best  method  of  installing  the  valve  spindle, 
as  the  spindle  remains  in  the  seat  when  the  valve  is  removed. 


292 


PUMPING  MACHINERY 


At  times,  however,  the  spindle  and  nut  are  combined  into  one 
piece,  Fig.  242,  and  in  placing  this  in  position  a  plug  wrench 
is  used. 

The  valve   disc    is   sometimes   backed   up   by  a  brass  cup 


FIG.  242. — Valve  Spindle. 


FIG.  243.— Valve  Barking. 


FIG.  244. — Metal  Disc  Valve. 


FIG.  245. — Cameron  Valves. 


into  which  it  fits  (Fig.  243).-  This  gives  more  stiffness  to  the 
valve. 

When  hot  water  is  used  brass  valves  replace  the  composition 
valves  which  are  intended  for  cold  water,  although  there  are 
special  compositions  used  for  hot  water.  Fig.  244  illustrates 
the  form  often  used  for  metal  valves. 

A  neat  arrangement  for  the  suction  and  discharge  valves 
is  made  by  the  Cameron  Company.  These  valves  (Fig.  245 ) 


DESIGN  OF  PARTS 


293 


are  placed  on  a  common  spindle  which  is  held  in  place  by  a  set 
screw  in  a  cap,  screwed  into  the  wall  of  the  valve  chest.  This 
spindle  also  serves  to  hold  the  valve  seats  in  place.  These 
valves  are  backed  with  metal  cups. 

The  Marsh  pump  is  equipped  with  metal  valves  the  spindles 


FIG.  246. — Marsh  Valves. 

of-  which  are  guided  by  holes  in  the  metal  at  the  center  of  the 
valve  seat.  This  center  is  so  formed  that  the  water  is  deflected 
under  the  valve  as  shown  in  Fig.  246.  This  central  part  is 
almost  as  large  as  the  cavity  of  the  valve  and  when  the  valves 
are  raised  the  distance  shown  in  the  figure,  there  is  no  tendency 


FIG.  247. — Rubber  Valve  with  Guard. 

for  them  to  lift  higher.  This  puts  them  into  a  position  where 
they  may  seat  quickly,  but  without  shock,  as  the  deflector  acts 
as  the  piston  of  a  dash  pot. 

When  rubber  valves  are  made  of  large  diameter  they  are 
designed  as  shown  in  Fig.  247.  The  guard  on  the  back  of  the 
valve  keeps  the  valve  from  opening  too  far.  The. holes  in  the 


294 


PUMPING  MACHINERY 


center  permit  the  water  to  act  on  the  back  of  the  disc  and  aid 
in  quick  closing.  The  method  of  supporting  the  spindle,  and 
thus  holding  the  seat  in  position,  is  clear  from  the  figure. 


FIG.  248.— Double  Ported  Valve. 

To  increase  the  opening  in  the  valve  discs  of  large  diameter 
they  are  made  multiported;  in  reality  they  become  a  series 
of  concentric  rings,  joined  by  radial  arms,  as  was  explained  in 


FIG.  249. — Riedler  Valve. 

Chapter  V.  Fig.  248  illustrates  such  a  valve  with  seat,  spindle, 
spring,  and  nut  where  one  ring  only  is  used.  The  form  of  this 
type  of  valve,  used  by  the  Allis-Chalmers  Co.  in  their  Riedler 
pumps,  is  best  shown  by  Fig.  249.  In  this  valve  a  leather 


DESIGN  OF  PARTS 


295 


washer  is  used  in  addition  to  the  conical-faced  ring  and  seat. 
The  leather  is  held  between  two  iron  rings.  This  leather  really 
makes  the  joint  when  the  pump  is  in  action.  The  valve  may 
be  raised  from  its  seat  by  the  water  pressure  when  the  arms 
CC  are  raised.  The  amount  of  motion  is  limited  by  the  nut 
D  on  the  spindle.  At  the  end  of  the  stroke  the  arms  CC  are 
driven  down  against  the  metal  sleeve  B  by  the  action  of  an 
eccentric.  This  presses  against  a  rubber  collar  A  and  forces 
the  valve  to  its  seat.  The  rubber  collar  gives  a  yielding 


FIG.  250. — Weighted  Valve. 

connection  to  care  for  the  possibility  of  solid  objects  getting 
beneath  the  valve.  Before  the  pump  reaches  the  other  end 
of  its  stroke  the  arms  CC  are  raised  so  that  the  valve  may 
open  as  soon  as  the  pressure  is  sufficient.  The  Riedler  valve 
was  first  employed  with  express  pumps. 

A  weighted  ring  valve  of  American  design  is  shown  in 
Fig.  250,  although  for  smaller  sizes  springs  could  be  used,  as 
shown  in  Fig.  251.  This  valve  (Fig.  250)  was  employed  on  the 
Lawrence  pump  to  replace  the  double-beat  valve  of  Fig.  252. 
Where  large  area  is  required,  double-beat  valves  (Figs.  252 


296 


PUMPING   MACHINERY 


and  253)  are  used.  Double-beat  valves  are  valves  containing 
two  seats,  as  at  A  and  B.  The  upper  seat  B  is  made  smaller 
than  that  at  A ,  and  hence  an  upward  pressure  from  below  causes 
the  valve  to  lift  from  its  seat  and  the  water  to  escape  by  these 
two  openings.  This  gives  a  large  area  and  by 
properly  arranging  the  size  of  each  seat,  a 
proper  amount  of  pressure  increase  can  be  had 
to  lift  the  valve  weight.  The  motion  of  the 
valve  is  limited  by  the  nut  on  the  spindle.  In 
Fig.  252  the  valve  is  open  to  its  full  extent. 
Although  these  valves  have  been  used  to  some 
FIG.  251.— Double  degree,  the  American  practice  is  to  use  a  num- 
ber of  small  valves,  and  where  sufficient  valve 
deck  area  is  not  available  valve  caps  (Fig.  254)  may  be  placed 
over  the  openings  used  for  double-beat  valves.  In  this  manner 
a  large  discharge  area  may  be  obtained.  The  valve  cap  or 
box  is  held  to  the  main  valve  deck  by  a  large  through  bolt. 
A  series  of  webs  are  employed  within  this  to  stiffen  it. 


FIG.  252. — Double  Beat  Valve. 


FIG.  253. — Double  Beat  Valve. 


The  valves  shown  in  Fig.  255  illustrate  the  method  used  for 
high-speed  pumps.  These  valves  have  been  recently  introduced 
by  Witting  of  England.  In  them,  two  brass  rings  are  placed 
around  each  opening.  The  rubber  rings  BB  force  the  brass 
rings  together  when  the  water  pressure  on  each  side  of  the 
brass  ring  is  the  same.  The  rings  touch  each  other,  and  are 


DESIGN  OF  PARTS 


297 


held  together  by  the  water  pressure  on  the  outside  as  well  as 
by  the  rubber.     When  the  internal  pressure  is  sufficient  to  force 


FIG.  254. — Valve  Box. 

these  apart,  the  brass  ring?  are  separated  and  this  give?  a 
series  of  large  openings  to  which  the  water  is  guided.  These 
valves  may  be  made  with  leather  facings,  and  helical  springs 
may  replace  the  rubber.  The  lower  figure  illustrates  a  method 


298 


MACHINERY 


mm 

FIG.  255.— Witting's  Metallic  Valves. 


FIG.  256. — Pressure  Valve. 


DESIGN  OF  PARTS 


299 


of  increasing  the  valve  area.  These  valves  are  so  constructed 
that  they  may  replace  other  valves,  the  cages  being  bolted 
to  the  valve  decks.  These  valves  are  somewhat  similar  to  the 


rm 


FIG.  257. — Pressure  Valves. 

valves  used  on  Merryweather's  fire  engine  of  1870,  in  which 
rubber  replaced  the  metal  rings. 

In  high-pressure  pumps  the  valves  are  made  deep  for  stiff- 
ness; the  springs  are  usually  made  heavier  on  the  discharge 
side  to  close  the  valve  quickly,  as  considerable  spring  pressure 
is  not  objectionable  on  this  side.  The  cap  over  the  valve 
and  the  walls  are  made  heavy  and  the  valve  should  be  access- 
ible. Fig.  256  illustrates  a  simple  form  of  valve,  as  used  with 


300 


PUMPING  MACHINERY 


the  Burnham  pump,  while  Fig.  257  shows  a  series  of  valves 
on  one  deck.  The  springs  are  held  in  place  by  a  ring.  This 
makes  it  possible  to  remove  the  cover  of  the  valve  chamber 
for  inspection  without  disturbing  the  springs.  In  these  two 
arrangements  the  springs  are  guided  by  spindles  on  the  caps, 
frame,  or  valve.  When  this  cannot  be  done  a  frame  is  formed 
on  top  of  the  valve  and  placed  around  the  valve  and  its  frame 
(Fig.  258).  Such  a  cage  may  be  held  down  by  a  ring  and  bolts 
(Fig.  259),  as  was  also  used  in  Fig.  257.  The  massive  construc- 
tion shown  in  the  last  four  figures  is  due  to  the  excessive 


FIG.  258. — Pressure  Valve  and  Cage. 


FIG.  259.— Valve  Pot. 


pressure  carried.  In  all  of  them,  it  is  seen  that  the  parts  may  be 
easily  removed  and  repaired,  and  quick  seating  of  the  valve 
is  possible  so  that  the  slip  may  be  cut  to  a  minimum. 

A  form  of  valve  used  by  German  pump  builders  is  known  as 
the  Gutermuth  valve. 

In  this  valve  (Fig.  260)  a  sheet  of  steel  or  bronze  is  bent 
into  the  form  of  a  spiral  which  forms  a  spring  for  the  end  of  the 
strip,  which  is  left  flat  to  form  the  valve.  A  pressure  from 
beneath  drives  the  plate  into  the  dotted  position,  and  the  water 
has  a  path  through  which  the  flow  occurs  almost  straight  from 
the  channel  leading  to  the  valve.  This  channel  is  not  arranged 


DESIGN  OF  PARTS 


301 


normal  to  the  valve  disc,  as  it  is  desired  to  continue  the  flow 
in  as  straight  a  line  as  possible. 


FIG.  260. — Gutermuth  Valve. 


JTL 


FIG.  261. — Borsig  Valve. 

The  Borsig  Co.,  of  Germany,  use  a  metal  valve  of  very 
simple  form.  It  is  made  of  thin  spring  metal  of  the  form  shown 
in  Fig.  261.  The  points  A  A  are  bolted  to  the  valve  deck, 


302 


PUMPING  MACHINERY 


and  the  arms  extending  to  the  outer  ring  form  the  springs  to 
seat  the  valve.  At  times  additional  helical  springs  are  placed 
on  the  back  of  this. 

The  discharge  air  chambers  for  pumps  are  made  in  various 


<Q          [j@5 


forms,  as  shown  in  Fig.  262.  For  small  pumps  the  first  form 
is  often  used  with  the  main  body  of  sheet  copper,  shaped  into  a 
conical  vessel.  The  other  forms  are  usually  made  of  cast  iron, 
and  are  placed  on  top  of  the  discharge  valve  chamber.  The 
second  form  is  well  arranged,  as  the  main  discharge  is  cared 
for  by  the  bend  in  the  pipe.  The  sizes  of  these  are  fixed  by 


DESIGN  OF  PARTS 


303 


considerations  given  in  Chapter  V,  and  the  thickness  is  deter- 
mined by  the  formulas  given  in  this  chapter.  As  was 
pointed  out  earlier,  the  size  of  the  air  chamber  will  depend  on 
the  details  of  an  installation,  so  that  a  standard  size  cannot 
be  fixed  for  stock  pumps  which  will  be  suitable  for  any 
purpose. 

The  air  from  the  chamber  is  gradually  absorbed  by  the 
water,  and  to  keep  the  chambers  charged  air  compressors  may 
be  used.  A  simple  device  (Fig.  263)  is  sometimes  applied. 
This  consists  of  piping  attached  to 
the  pump  cylinder  and  air  chamber 
in  which  are  placed  two  check  valves 
and  one  gate  valve.  The  check 
valves  open  toward  the  air  chamber. 
When  the  valve  is  opened  air  is 
drawn  into  the  vertical  line  on  the 
suction  stroke  of  the  pump,  and  on 
the  discharge  stroke  this  is  com- 
pressed and  sent  over  into  the  air 
chamber,  as  the  air  cannot  escape  to 
the  atmosphere  and  the  pressure  in 
the  cylinder  is  greater  than  the  pres- 
sure in  the  air  chamber.  On  each 
stroke  there  is  a  slight  discharge  of 
air  into  the  chamber,  and  after  suf- 
ficient air  is  obtained  the  stop  valve  is  shut  off  and  the 
operation  of  this  air  pump  ceases. 

The  air  chambers  on  the  suction  side  of  the  pump  are 
important  adjuncts,  as  was  seen  in  Chapter  V.  They  should 
be  placed  so  that  they  can  receive  and  care  for  the  water  which 
it  is  necessary  to  store  for  the  variation  in  the  suction  of  the 
pump.  The  air  chambers  of  Fig.  264  are  all  placed  at  the  end 
of  a  suction  line,  while  the  water  is  taken  to  the  valve  chest 
from  the  side  of  this  line.  The  various  arrangements  illustrate 
how  the  air  chamber  can  be  placed  when  different  conditions 
of  piping  exist.  In  all  cases  the  water  is  flowing  directly  into 
the  air  chamber. 


FIG.  263. — Air  Chamber 
Pump. 


301 


PUMPING  MACHINERY 


Fig.  265  shows  two  air  chambers;    the  left-hand  being  a 
simple  one,  made  of  pipe  fittings,  and  of  proper  design,  while 

Ift 


the  right  is  improperly  arranged,  since  the  water,  in  passing 
at  right  angles  to  the  air  chamber,  must  exert  some  force  on  the 
valves  beyond  before  the  air  in  the  chamber  is  compressed. 


DESIGN  OF  PARTS 


305 


The  left-hand  chamber  discharges  from  the  end  B  at  a  varying 
rate,  while  the  supply  at  A  is  uniform.  The  excess  or  deficiency 
is  made  up  from  the  water  which  passes  into  the  space  around 
the  upper  pipe.  This  flow  goes  naturally  into  the  space,  as  the 
stream  enlarges  suddenly.  To  cut  down  the  loss  due  to  sudden 
enlargement  and  yet  get  the  advantage  of  it,  the  suction  air 
chamber  could  be  built  as  shown  in  Fig.  266,  and  in  this  manner 


FIG.  265. — Suction  Air  Chambers. 


FIG.  266. — Suction  Air  Chamber. 


the  losses  of  sudden  enlargement  and  contraction  could  be 
reduced. 

To  protect  the  pump  valves  against  leaves,  twigs,  and  other 
foreign  particles  which  might  lodge  beneath  the  valves  and 
prevent  their  proper  action,  strainers  (Figs.  267,  268,  269  and 
270)  are  used.  These  are  made  of  wires,  grills,  or  drilled  plates. 
They  should  always  be  so  constructed  that  they  may  be  cleaned 
easily.  At  times  the  strainer  is  combined  with  a  foot  valve, 
as  in  Figs.  269  and  270.  A  foot  valve  is  a  large  valve  or  series 
of  small  valves  attached  to  a  valve  deck  as  shown  in  the  figures, 
and  placed  on  the  end  of  the  suction  main.  The  purpose  of 


PUMPING  MACHINERY 


the  foot  valve  is  to  hold  the  water  in  the  suction  main  when  the 
pump  is  shut  down,  and  to  keep  the  line  filled  in  case  it  is  very 
long.  The  large  foot  valve  may  be  of  the  lift  form,  as  shown 


FIG.  267. — Suction  Strainer. 


IT 


FIG.  268. — Strainer. 


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FIG.  270. — Foot  Valve  and  Strainer. 


The 


in  Fig.  270,  or  it  may  be  of  the  form  shown  in  Fig.  271. 
type  shown  in  Fig.  268  is  arranged  as  shown  in  Fig.  271. 

The  suction  pipe  should  be  of  large  diameter,  as  straight 
and  as  short  as  possible,  and  air  tight.  The  smallest  leak  is 
apt  to  destroy  the  proper  action  of  the  pump.  It  is  well  to 
test  these  pipes  under  pressure  before  covering  them.  The 


DESIGN   OF  PARTS 


307 


pipe  must  be  laid  on  a  uniform  grade  to  prevent  air  pockets. 
Everything  should  be  done  to  cut  down  the  losses,  as  the  total 
loss  with  the  lift  must  be  kept  less  than  34  feet. 

With  these  details  in  mind,  the  design  will  now  be  con- 
tinued. The  number  of  revolutions,  the  size  of  the  pump, 
the  size  and  number  of  the  valves,  the  sizes  of  the  pipe  lines, 


FIG.  271. — Foot  Valves. 

and  the  size  of  the  air  chambers  are  determined  as  in  Chapter 
V.  From  these,  the  pressure  in  the  water  cylinder  can  be 
found. 

DESIGN  OF  CYLINDER 

In  the  water  cylinder  (Fig.  272)  the  first  quantity  to  be 
determined  is  the  thickness  of  the  walls. 

When  the  cylinder  walls  are  cylindrical  and  the  surface 
is  not  subject  to  wear,  the  formula: 

pd  =  2iS  or  its  equivalent  may  be  used; 
p=the  water  pressure  in  Ibs.  per  sq.in.; 
d  =  diameter  to  center  of  walls  in  inches; 
t  =  thickness  in  inches; 
5=  allowable  stress  in  Ibs.  per  sq.in. 

This  formula  is  for  thin  cylinders;    for  thick  cylinders  used 


308 


PUMPING  MACHINERY 


with  heavy  pressures  the  formulae  of  Banow  and  Lam6  are 
employed. 

Before  proceeding  with  these  formulae  it  will  be  well  to 
note  the  table  (page  309)  of  the  strengths  and  other  properties 
of  the  materials  to  be  used  in  design  together  with  the  safety 
factors  required.  These  values  have  been  taken  from  standard 


FIG.  272.  —  Pump  Section. 

texts  on  machine  design  as  well  as  from  engineering  handbooks 
and  current  literature. 

The  formula  above  where  cast  iron  is  used  as  the  cylinder 
wall  is  put  into  the  form 


. 

5000 

When  the  cylinder  is  bored  for  the  reception  of  the  piston 
and  allowance  is  made  for  reboring,  this  equation  becomes 


.. 

5000 

In  making  the  head  as  shown  in  the  Fig.  272,  the  radius  of 
the  spherical  portion  should  be  made  equal  to  the  diameter 


DESIGN  OF  PARTS 


309 


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Wrought  iron  .  . 

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§  |  i  i 

310  PUMPING  MACHINERY 

of  the  cylinder,  and  then  the  thickness  should  just  equal  that  of 
the  walls. 

If  the  head  has  to  be  made  flat  the  formula  for  the  thickness 
when  well  supported  by  bolts  is 


for  a  round  plate  (d  =  diameter),  while 


* =0.350  \l 


'S 

holds  for  a  square  plate  (a=  side  of  square),  and 

ab 


for  a  rectangular  plate  of  length  a  and  width  b. 

For  manhole   or   handhole   covers  which   are    of   elliptical 
form,  the  thickness  is  fixed  by  the  formula 

ab 


In  all  cases  the  endeavor  should  be  made  to  eliminate  flat 
places,  and  when  such  must  exist  they  must  be  designed  for 
thickness  as  outlined  above.  At  times  ribs  are  used  to  stiffen 
flat  places,  and  at  times  stay  bolts  are  used  crossing  from  side 
to  side. 

The  thickness  of  heavy-pressure  pump  cylinders  is  deter- 
mined by  either  one  of  two  formulae.  Barlow's: 

f      Pr* 
~S-p' 
or  Lamp's: 


=  S' 


t=  thickness  in  inches; 
p=  hydrostatic  pressure  in  Ibs.  per  sq.in.; 
TI  =  the  outside  radius  of  cylinder,  in  inches; 
y2=the  inside  radius  of  cylinder  in  inches; 
S=maximum  stress  in  Ibs.  per  sq.in. 


DESIGN  OF  PARTS 


311 


In  designing  the  cylinders  for  hydraulic  presses  or  pressure 
pumps  the  metal  thickness  should  be  uniform,  as  thicker  parts 

of  castings  cool  after  the  other  parts 
have  solidified.  There  is  danger,  there- 
fore, that  cooling  strains  will  be  set  up 
and  cracks  may  develop. 

There  should  be  no  sudden  changes 
in  the  thickness  of  the  section  and  all 
corners  must  be  well  filleted.  These 
points  should  be  observed  in  all  cylinder 
design  or  in  any  other  castings. 

When  the  core  of  the  heavy  cyl- 
inder castings  has  to  be  supported  at  each  end  a  hole  must  be 
left  at  the  bottom  of  the  cylinder,  and  when  such  is  the  case 
it  is  closed  by  driving  in  a  plug  or  by  using  a  packed  joint 
(Fig.  274)  in  which  a  hydraulic  packing  of  some  form  is 
used. 


FIG.  273. — Pressure 
Cylinder. 


FIG.  274. — Core  Hole  Cap. 


FIG.  275. — Twin-Cylinder  Casting. 


When  two  hydraulic  cylinders  are  to  be  used  in  one  casting, 
there  must  be  sufficient  metal  where  the  two  cylinders  come 
together  to  support  the  total  load  in  each  cylinder.  The  upper 
part  of  Fig.  275  shows  the  incorrect  method,  while  in  the  case 
below  there  is  sufficient  metal  to  carry  the  load. 

When  there  are  to  be  openings  made  into  cylinder  walls 


312 


PUMPING  MACHINERY 


for  pipes,  manholes,  or  other  purposes,  the  edge  of  the  hole 
should  be  reinforced  by  a  ring  or  boss,  and  the  amount  of  metal 
added  should  equal  or  should  exceed  the  amount  of  metal 
removed  in  the  section,  as  shown  in  Fig.  276.  In  this  figure 
the  amount  of  metal  cut  away  is  dt  square  inches,  and  this 


*      t 


1 


FIG.  276. — Branch  Connection  Reinforcement. 

should  be  equal  to  or  less  than  (a—d}h.  These  openings 
should  be  avoided,  but  when  necessary  in  every  case  the  rein- 
forcement should  be  made. 


FLANGES 

The  flanges  to  be  used  on  cylinders,  heads,  manhole  covers, 
and  covers  for  other  parts  of  the  pumps  are  in  general  made 
equal  to  3  diameters  of  bolt  in  width;  this  gives  i^d  from 
the  center  of  bolt  to  edge,  a  proportion  which  is  common  in  all 
design  work.  Until  the  size  of  the  bolts  is  determined  this 
cannot  be  fixed.  It  is  best,  therefore,  to  assume  the  diameter 
of  bolt,  and  design  the  number  to  be  used.  The  bolts  should 
be  taken  as  large  as  possible.  Small  bolts  are  apt  to  be  twisted 
off.  It  is  well  to  use  nothing  less  than  f -inch  bolts  except  when 
very  small  parts  are  used.  On  cylinders  from  8  to  15  inches, 
J-mch  bolts  should  be  assumed;  from  15  to  24  inches,  i  inch; 
and  for  larger  sizes,  the  diameter  should  be  more  than 
i  inch. 

The  flange  width  is  now  made  equal  to  3^. 

The  thickness  of  the  flange,  t" ,  is  made  1.2  to  1.4^;  where 
t=  thickness  of  cylinder  walls. 

The  bolts    support  the    pressure  on   the   area   A   of  the 


DESIGN   OF  PARTS  313- 


cover  exposed  to  the  water  pressure.     The  number  n  is  given 
by 


a  =  area  of  bolts  at  root  of  thread  found  in  table  of 

bolts  in  sq.in.; 

5=  allowable  stress  in  Ibs.  per  sq.in.; 
A  =area  supported  in  sq.in.   (usually  to  the  inner  edge 
of  cover,  although  if  packing  is  not  tight  this  area 
may  be  increased); 
p  =  water  pressure  in  Ibs.  per  sq.in. 

After  finding  the  number  of  bolts,  the  pitch  of  the  bolts 
should  be  found,  and  if  this  is  greater  than 


where  t"  is  the  thickness  of  the  flange,  more  bolts  must  be 
used.  The  quantity  above  is  the  maximum  allowable  pitch  for 
a  tight  joint. 

WATER  PISTONS 

The  water  piston  is  made  up  of  various  parts,  but  it  may 
be  considered  as  a  flat  plate  supported  at  the  center  and  loaded 
uniformly.  In  this  case  the  thickness  would  be  about  two- 
thirds  the  thickness  of  the  cylinder  head.  If  the  piston  is 
cored,  as  in  Fig.  297,  Chapter  VIII,  this  thickness  is  made  less, 
as  the  casting  acts  like  a  deep  beam.  The  thickness  of  each 
part  could  be  made  less  than  one-third  the  thickness  of  the 
head  but  this  amount  may  be  used.  If  the  two  parts  are 
separate,  as  in  Fig.  199,  then  each  should  be  of  the  full 
thickness. 

Piston  design  is  in  most  cases  empirical  and  the  proportional 


number  is  -  -  for  cast  iron  or  bronze.     This  will  be  con- 
100 

sidered  in  Chapter  VIII,  and  until  that  place  the  proportions 
will  not  be  given. 


314  PUMPING  MACHINERY 

The  packing  used  on  water  cylinders  is  often  made  of 
square-plaited  flax,  hemp  or  cotton,  with  a  rubber  core,  indurated 
with  rubber  or  alternate  layers  of  rubber  and  cotton.  The 
sizes  are  J-XJ  inch,  f  Xf  inch,  JXJ  inch,  etc.  In  putting  this 
in,  the  packing  should  be  cut  so  that  the  ends  do  not  touch  by 
about  the  thickness  of  the  packing.  This  packing  swells  on 
being  wet.  The  depth  of  the  packing  is  0.4^  on  Fig.  199. 

PISTON  RODS 

The  piston  rod  is  designed  to  take  compression  only  in  a 
single-acting  pump,  and  to  take  compression  and  tension  in  a 
double-acting  pump.  In  many  cases  these  rods  are  of  steel, 
but  in  some  small  pumps  where  corrosion  is  to  be  avoided, 
phospor  bronze,  Muntz  metal,  or  some  other  form  of  non- 
rusting  metal  is  used. 

Tobin  bronze  is  an  alloy  of  copper  (59%),  zinc  (38.4%), 
tin  (2.16%).  Iron  (0.11%)  and  lead  (0.31%)  are  sometimes 
found  in  the  mixtures.  It  has  a  tensile  strength  as  high  as 
75,000  when  rolled,  and  at  times  100,000  pounds  per  square 
inch,  has  been  obtained  after  cold  rolling.  At  a  cherry  red 
this  bronze  can  be  stamped  and  forged. 

Muntz  metal  is  an  alloy  of  3  parts  of  copper  with  two  of 
zinc.  It  has  a  strength  of  about  50,000  pounds  per  square  inch. 

The  piston  rod  is  treated  as  a  column  in  all  cases,  as  it  is 
the  compressive  load  which  fixes  the  size. 

The  figures  of  this  chapter  illustrate  the  method  of  attach- 
ment of  the  piston  rod.  It  will  be  seen  that  the  end  of  the 
rod  has  a  taper  extending  from  a  shoulder  or  collar  to  a  cylin- 
drical-threaded portion.  The  threaded  portion  takes  plain  ten- 
sion; the  shoulder  and  the  tapered  portion  take  the  com- 
pression. It  is  well  in  forming  the  shoulder  or  collar  to  fillet 
the  corner  instead  of  leaving  it  absolutely  square,  as  this  fillet 
increases  the  strength,  although  it  is  not  so  easy  to  get  a  fit 
of  the  rod  to  a  rounded  corner. 

The  load. on  the  rod  is 


DESIGN  OF  PARTS  315 

The  area  at  the  root  of  the  thread  is 

_P 

tt  —  c*  • 

The  diameter  of  this  thread  is  found  from  a  table  of  standard 
threads  for  the  area  a. 

The  rod  is  designed  by  a  column  formula: 

P  Sc 

,  Rankin's  Formula; 


p  i 

—-jz=Sc  —  K'—,  Straight-Line  Formula; 

TtCl  Y 

T 

P  =  total  load  on  one  rod  in  Ibs; 
d  =  diameter  of  rod  in  inches; 

d 

r  =  radius  of  gyration  =-; 

5C=  allowable  working  stress  in  Ibs.  per  sq.in.; 
/=  greatest  length  of  rod  between  supports  in  inches; 

K=  -  --  for  steel;  round  ends; 
6250 

K'=  284  for  steel;  round  ends. 

These  equations  must  be  solved  by  trial  by  first  getting  the 

p 
value  of  d  from  —  ^  =  5C,  and  then  using  this  to  find  r\   solve 


4 
for  d  after  the  approximate  value  of  r  has  been  inserted. 

The  shoulder  used  on  the  rod  should  be  j-  to  J  inch  on  small 
rod,  while  \  inch  may  be  used  if  a  collar  is  formed.  The  taper 
part  of  the  rod  should  reduce  the  section  to  the  diameter  at  the 
thread.  A  taper  such  that  the  diameter  is  reduced  3  inches  in 
a  toot  is  a  good  one  to  use. 

STUFFING  BOXES 

The  stuffing-box  design  has  been  given  earlier  in  this 
chapter. 


316  PUMPING  MACHINERY 

VALVES  AND  SPRINGS 

The  proportion  of  valves  will  be  seen  from  the  various 
figures.  The  diameter  of  the  valves  has  been  discussed  in 
Chapter  V.  The  spring  design  will  now  be  considered.  The 
formulae  for  springs  are: 


<•* 


where      5  =safe  shearing  strength  in  Ibs.  per  sq.in.; 
P=loadinlbs.; 
n  =  total  number  of  coils;     Jj  • 
d  =  diameter  of  wire  in  inches; 
£=torsional  modulus  of  elasticity; 
X=  compression  in  inches. 

From  formula  (a)  the  radius  of  coil  may  be  found  to  carry 
the  load  for  an  assumed  diameter  and  stress  for  the  spring 
wire.  With  this  radius  the  number  n  may  be  found  to  give 
the  necessary  compression  with  a  given  change  of  load. 


The  pitch  of  the  helix  is  found  by  assuming  the  height 
allowable,  .and  putting  n  turns  in  this  height.  If  this  result 
is  not  thought  proper  a  different  diameter  of  wire  may  be  taken; 
a  new  r  and  n,  found.  This  should  be  tried  until  one  of  the 
proper  form  is  found.  Although  r  and  n  are  definite  and  would 
give  the  desired  result,  they  do  not  always  fit  the  other  parts 
of  the  design. 

The  following  proportions  for  valves  are  used  in  practice, 
according  to  Barr: 


DESIGN  OF  PARTS  317 

VALVE  Discs 

Diameter.  Thickness.  Hole. 

2  "  f"  1" 

21"          A"  r 

3  "  r  ft" 
3i"  f"  f" 

4  "  f "  ft" 
4i"  I"  ft" 

r>        »  •  3"  13" 

5  4  T5 

Taper  of  thread  on  brass  valve  seat  to  fit  valve  deck  i  inch  per 
foot.  Valve  stem  ^  inch  less  than  diameter  of  hole.  Plate  on 
top  of  valve  ^  inch  thick,  and  of  diameter  equal  to  three-fourths 
diameter  of  disc.  Plate  ^  inch  thick  for  valve?  4^  inches  and 
over.  Springs  are  made  of  the  following  sizes: 

Diameter  of  Valve.  Size  of  Wire. 

2  "  NO.  12 
2j"  NO.  12 

3  "  No.  10 
3i"  No.  10 

4  "  No.    8 
4i"  No."  8 

Springs  are  usually  coiled  with  a  diameter  of  one-half  the  valve 
diameter,  and  with  five  turns  they  have  sufficient  elasticity. 

Fig.  277  illustrates  the  method  of  installing  a  small  modern 
pump  and  applies  the  general  principles  outlined  in  these  two 
chapters.  The  foot  valve  A  is  applied  at  the  lower  end  of  the 
suction  pipe.  The  suction  air  chamber  B  is  applied  so  as  to 
receive  directly  the  impulse  from  the  water.  The  strainer  at 
C  is  conveniently  placed  for  cleaning.  A  check  valve  D  per- 
mits one  to  relieve  the  valve  chamber  of  pressure  when  necessary 
to  examine  the  valves  without  draining  the  discharge  main. 
The  priming  pipe  E  is  used  to  prime  the  pumps.  By  opening 
the  waste-drain  pipe  F  and  the  primer  E  the  pump  is  filled 
with  water,  the  air  being  driven  out.  The  waste  pipe  may  be 
opened  in  starting  the  pump  so  that  air  may  be  removed  before 
the  full  pressure  is  exerted.  Such  an  arrangement  may  be 


318 


PUMPING  MACHINERY 


necessary  in  starting  a  compound  pump  where  pressure  is  needed 
on  both  pistons  to  start  the  pump  against  full  head.  In  such 
a  case  the  pump  could  not  start  unless  live  steam  were  by-passed 
into  the  low  pressure  cylinder.  Another  method  sometimes 
used,  which  is  the  equivalent  of  this,  is  to  connect  the  two 
ends  of  the  water  cylinder  by  a  cross  connection  to  eliminate 
water  pressure.  After  both  cylinders  have  received  the  proper 


FIG.  277. — Arrangement  of  Pump. 

steam  on  finishing  a  few  strokes  the  waste  or  the  connection 
between  the  ends  is  closed,  and  the  pump  discharges  against 
the  working  head. 

The  figure  illustrates  large-size  suction  pipes  and  short, 
direct  connections,  so  necessary  for  proper  action.  Care  must 
be  taken  in  making  up  these  joints  to  have  them  air  tight. 

The  following  quotations  from  the  catalogue  of  the  Snow 
Pump  Works  are  valuable,  although  many  of  them  have  been 
given  previously  in  this  chapter: 


DESIGN  OF  PARTS  319 

"  INFORMATION  CONCERNING  PIPE  CONNECTIONS 

"  Fig.  277  shows  in  a  general  way  the  proper  method  of 
piping  a  pump.  Faulty  connections  are  generally  the  cause 
of  the  improper  action  of  a  pump,  and  great  care  should,  there- 
fore, be  taken  to  have  everything  right  before  starting.  To 
accomplish  this,  note  carefully  and  understand  thoroughly 
the  following: 

"  Be  sure  that  the  quantity  of  water  you  desire  to  pump 
is  available  and  that  your  pump  is  within  easy  reach  of  it 
when  it  is  at  its  lowest  level. 

"  Locate  your  pump  as  near  the  source  of  suction  supply, 
both  vertically  and  horizontally,  as  is  possible  or  convenient; 
but  never  place  it  in  such  a  location  that  the  sum  of  the  follow- 
ing three  items  would  exceed  a  total  of  26  feet: 

"  i.  Height  in  feet  from  the  discharge  valves  of  the  pump 
to  the  lowest  level  of  the  surface  of  the  suction  water. 

"  2.  Total  frictional  loss  in  suction  pipe  in  feet  head. 

"3.  Total  frictional  loss  in  feet  head,  due  to  elbows  and 
tees  (assumed  as  being  the  equivalent  of  the  frictional  loss  due 
to  100  feet  of  same  size  of  pipe,  for  each  elbow  or  tee). 

"  EXAMPLE. — Would  a  pump  having  an  easy  capacity  of 
750  U.  S.  gallons  per  minute  operate  satisfactorily  at  this 
capacity  if  the  height  from  the  delivery  valves  to  the  surface 
of  the  suction  supply  was  20  feet,  the  suction  pipe  8  inches 
diameter,  and  800  feet  long,  and  having  two  8-inch  standard 
elbows? 

"ANSWER. — No.     (Ascertained  as  follows): 

"  Height  from  delivery  valves  to  surface  of  water. . .  =  20.  ft. 
"  Total  friction  in  800  ft.of  8-in.pipe  (  =  .53  x8  X2.3I)  =  9.79  ft. 
"Total  friction  in  two  8-in.elbows  (  =  .53x2.31X2)=  2.45  ft. 


Total 32.24  ft. 

"The  sum  of  these  three  items  is  in  this  case  about  32  feet, 
or  6  feet  greater  than  the  26  feet  above  mentioned.  Therefore, 
the  pump  should  be  lowered  6  feet,  or  the  frictional  resistance 


320  PUMPING   MACHINERY 

in  the  suction  pipe  reduced  by  about  6  feet,  by  increasing  the 
size  of  the  suction  pipe. 

"  Lay  your  suction  pipe  so  that  it  slopes  away  from  the 
pump  gradually.  A  suction  pipe  should  have  no  air  pockets 
in  its  entire  length,  but  should  be  so  laid  that  if  air  be  admitted 
to  it,  near  the  intake  end,  with  the  pump  standing  still,  the 
air  would  rise  to  the  pump  or  suction  air  chamber,  and  not  be 
pocketed  in  some  high  part  of  the  suction  pipe.  A  slope  of 
i  per  cent  (i  foot  drop  in  100  feet  of  length)  will  be  found 
very  satisfactory. 

"  Be  sure  that  your  suction  piping  is  absolutely  tight,  for 
a  very  small  air  leak  will  cause  a  pump  to  work  improperly. 
The  suction  pipe  should  be  tested  with  about  20  pounds  water 
pressure  after  it  has  been  laid  and  before  it  is  covered.  If  the 
test  shows  up  a  leak,  fix  it;  it  is  not  '  good  enough.' 

"  Keep  the  end  of  your  suction  pipe  well  under  water. 
It  should  never  have  less  than  3  feet  above  it  and  6  or  8  feet 
would  be  much  batter. 

"  When  you  take  suction  from  a  tank  into  which  the  returns 
from  hydraulic  elevators  or  presses  empty,  take  care  that  the 
returns  enter  the  tank  as  far  away  from  the  suction-pipe  opening 
as  possible,  for  the  pump  is  liable  to  get  air  if  the  returns  empty 
near  the  suction  opening. 

"  If  two  or  more  pumps  draw  from  the  same  suction  pipe, 
or  if  water  comes  to  the  pump  under  a  head,  a  gate  valve  should 
be  placed  on  the  suction  pipe  of  each  pump,  to  enable  you  to 
open  up  any  one  pump  cylinder  for  repairs  or  examination 
without  interfering  with  the  operation  of  the  other  pumps. 
We  recommend  on  larger  sizes,  when  practicable  and  when 
not  too  costly,  that  each  pump  have  a  separate  individual 
suction  line  entirely  independent  of  the  suction  line  of  any 
other  pump. 

"  A  suction  air  chamber  will  be  found  desirable  in  all  cases, 
and  indispensable  in  cases  where  the  sum  of  the  three  items 
referred  to  in  a  previous  paragraph  exceeds  10  feet,  or  when 
the  suction  pipe  is  long. 

"  A  foot  valve  is  desirable  in  all  cases  (except  when  suction 


DESIGN   OF  PARTS  321 

water  comes  to  the  pump  under  a  head)  and  indispensable 
when  the  suction  lift  exceeds  10  feet.  By  its  use  the  pump 
and  suction  pipe  are  kept  primed  when  the  pump  is  shut  down, 
and  permits  of  easily  priming  the  pump  and  suction  pipe  if 
purposely  emptied,  thus  enabling  the  pump  to  be  easily  started 
at  any  time. 

"  In  all  cases  where  the  water  contains  sticks,  weeds,  rags, 
or  other  rubbish  a  strainer  should  be  used  on  the  suction  pipe, 
to  prevent  them  from  getting  into  the  pump  and  clogging 
valves  and  passages.  If  a  foot  valve  is  used,  a  strainer  placed 
outside  the  foot  valve  is  the  best;  but  if  no  foot  valve  is  used, 
a  box  strainer,  placed  near  the  pump,  and  so  designed. that  by 
removing  the  strainer  cover  all  accumulations  can  be  removed, 
will  be  found  the  most  desirable.  Keep  the  strainer  clear  from 
accumulation  of  rubbish. 

"  When  a  foot  valve  is  used,  a  drain  valve  should  be  placed 
near  the  surface  of  the  water,  to  enable  the  suction  pipe  to  be 
drained  when  desired. 

"  A  relief  valve,  set  to  blow  at  about  20  pounds  pressure, 
should  also  be  placed  on  the  suction  pipe  near  the  pump,  to 
prevent  delivery  pressure,  if  over  50  pounds,  from  accumu- 
lating in  the  suction  chamber  of  the  pump  or  the  suction 
pipe.  This  does  not  cost  much  and  may  sometime  save  you 
the  cost  of  replacing  a  broken  pump  cylinder  or  foot  valve, 
due  to  carelessness. 

"  In  cases  where  the  pump  gets  its  supply  directly  from 
driven  wells,  it  will  be  found  most  satisfactory  to  place  a  large 
tank  on  end  in  some  part  of  the  pump  house,  and  to  run  the 
pipes  from  the  wells  into  the  side  of  the  tank  not  far  from  the 
bottom.  Take  the  suction  for  your  pump  from  the  bottom 
of  the  tank,  and  lead  a  small  pipe  from  the  top  of  the  tank 
to  an  air  pump  provided  for  the  purpose,  or  to  the  condenser 
of  the  main  pump,  if  its  air  pump  is  large  enough.  A  gauge 
glass  on  the  side  of  the  tank  will  show  the  level  of-  the  water 
in  same,  and  by  opening  the  valve  on  the  small  air  pipe  a  certain 
amount  of  the  air  may  be  removed  from  the  tank  and  the 
pump  will  then  get  no  air,  but  pump  its  full  displacement  of 


322  *     PUMPING  MACHINERY 

water.  The  air  may  be  extracted  automatically  from  the 
tank  by  means  of  a  float  arrangement. 

"  A  check  valve  D  on  the  discharge  pipe  will  be  found  very 
convenient. 

"  A  gate  valve  should  always  be  placed  on  the  discharge 
pipe  outside  of  the  check  valve. 

"  A  priming  pipe  should  always  be  connected  from  the 
discharge  pipe,  outside  of  the  gate,  to  the  suction  pipe,  if  a 
foot  valve  is  used.  This  will  enable  the  pump  cylinders  and 
suction  piping  to  be  primed,  if  empty,  before  starting.  If  you 
have  no  suitable  relief  valve  on  the  suction  pipe,  be  very  careful, 
in  priming  with  this  pipe,  that  you  do  not  let  delivery  pressure 
accumulate  in  the  suction  pipe.  This  will  be  prevented  by 
having  the  starting  waste  valve  F  open  before  you  open  the 
priming  pipe  valve  E.  This  should  always  be  open  before  start- 
ing your  pump  (whether  you  have  a  foot  valve  or  not),  as  by  this 
means  the  pump  is  enabled  to  discharge  the  air  from  the  pump 
cylinders  and  suction  pipe  through  this  starting  valve  against 
a  light  pressure.  As  soon  as  water  is  discharged  through  the 
starting  valve,  shut  it  and  open  your  steam  throttle  valve, 
and  the  pump  will  then  discharge  through  the  discharge  main, 
opening  the  check  valve  automatically.  If  you  have  a  foot 
valve  or  a  gate  on  your  suction  pipe,  and  no  relief  valve,  be 
careful  to  open  the  starting  valve  at  the  instant  you  shut  the 
pump  down  and  leave  it  open  until  after  you  have  started 
again,  as  by  so  doing  you  prevent  the  possibility  of  pressure 
accumulating  in  the  suction.  The  pet  cock  in  the  force  chamber 
of  small-size  pumps  is  intended  to  be  used  in  the  same  manner 
as  the  starting  valve  above  referred  to. 

"  When  shutting  the  pump  down  in  late  fall,  winter,  or 
early  spring,  be  sure  and  open  all  steam  and  pump-end 
drains,  and  leave  them  open,  otherwise  your  steam  or  pump 
cylinders  or  other  parts  may  be  cracked,  due  to  water  freez- 
ing inside  of  them,  and  you  may  be  forced  to  purchase 
practically  a  new  pump.  A  little  care  will  prevent  this  occur- 
rence. 

"  Do  not  try  to  raise  hot  water,  crude  petroleum,  or  any 


DESIGN  OF  PARTS  323 

thick  liquids  by  suction,  but  wherever  possible  have  the  liquid 
flow  to  the  pump  under  a  suction  head. 

"  Keep  the  steam  cylinders  and  valve  motion  of  your  pump 
well  lubricated.  Oil  is  cheaper  than  repair  parts. 

"  Do  not  pack  the  stuffing  boxes  too  tightly,  and  do  not 
let  the  packing  stay  in  until  it  gets  hard  and  cuts  the  piston 
rods  or  plungers.  Renew  it  sufficiently  often  to  keep  it  soft 
and  pliable.  If  the  pump  runs  badly,  make  sure  that  the 
pump  valves,  packed  pistons,  or  plungers,  and  suction  and 
discharge  connections  are  all  right  before  examining  the  steam 
end." 


CHAPTER  VII 


DYNAMICS  OF  STEAM  END 

IN  the  design  of  a  pump  the  variation  of  pressure  has  been 
found  together  with  the  size  of  the  pump  cylinders,  and  it  is 
now  necessary  to  find  the  size  of  the  steam  cylinders  to  operate 
the  water  end,  the  size  of  the  electric  motor  to  drive  the  pump, 
or  the  size  of  the  water  wheel  which  is  required. 

The  curves  constructed  in  Figs.  197  and  198  can  be  com- 
bined to  give  the  indicator  card  of  the  head  end  of  the  pump. 
The  crank-end  card  of  the  water  end,  if  the  pump  were  double 

n  acting,  would  be  found  in  a 


Chan  her  I  ressu  ;e_222|ft 


Atmosph 


to  eric  Line 


|. 


FIG.  278.— Pump  Card. 


similar  manner.  From  this 
combined  card  (Fig.  278), 
which  is  the  indicator  card, 
the  power  to  be  given  to  the 
water  may  be  computed. 

Assume  that  there  are 
three  single-acting  pumps  giv- 
ing cards  similar  to  Fig.  278 
connected  together,  the  dis- 
charge pipes  of  each  cylinder  being  18  inches  in  diameter,  and 
continued  separately  to  the  height  considered  in  Chapter  V. 
The  area  of  the  card  then  would  give  the  work  done  on  the 
water.  The  mean  height  of -the  card  is  235.2  feet  or.  102.1 
pounds  per  square  inch.  This  includes  all  friction  of  the  water, 
and  would  give  the  power  required  if  an  allowance  is  made  for 
the  mechanical  efficiency  of  the  pump. 
The  water  indicated  horse  power  is 


I.H.P.  - 


3PLAN 
33,000  ! 


324 


DYNAMICS  OF  STEAM  END  325 

where       P=mean  effective  pressure  in  Ibs.  per  sq.in.; 
L=  stroke  in  feet; 
A  =area  of  piston  in  sq.in.; 
N  =  number  of  revolutions  per  minute; 
•    IHp=3Xio2.iXffX3i4X6o 

33,OOO 

The  mechanical  efficiency  of  the  water  ends  of  pumps  will 
vary  with  the  construction  and  condition  of  packing.  With 
a  plunger  pump  an  efficiency  of  97  per  cent  is  not  uncommon; 
and  this  will  be  taken  as  the  mechanical  efficiency  in  the  problem 
studied.  With  inside-packed  pumps  the  friction  may  increase 
so  that  for  such  pumps  92  per  cent  will  be  used.  These  values 
will  vary,  and  if  the  glands  or  packing  rings  on  pistons  are 
tightened  too  much  these  values  will  be  smaller. 

The  mechanical  efficiency  of  the  complete  steam-driven 
unit,  including  the  power  for  the  air  pump  of  the  condenser 
on  large  pumps  is  about  95  per  cent,  while  90  per  cent  or  even 
80  per  cent  should  be  used  for  small  pumps.  With  a  gear- 
driven  pump  the  gears  will  absorb  about  5  per  cent  of  the 
power  per  pair,  so  that  with  a  back  gear  the  loss  of  the  gearing 
amounts  to  10  per  cent.  The  efficiency  of  engines  and  electric 
motors  may  be  taken  as  90  per  cent  to  95  per  cent,  and  the 
efficiency  of  water  turbines  as  80  per  cent.  Using  the  figures 
above,  the  indicated  horse  power  of  the  engine  for  driving,  the 
electrical  power  and  the  water  horse  power  applied  to  the 
wheel  will  be  respectively: 


W.H.P.=        437'2    Q   =626  W.H.P. 

.97  x.  90  x.  80 

To  carry  further  the  design  of  the  steam  end  for  the  purpose 
of  finding  the  size  of  fly  wheels  or  other  parts  it  will  be  assumed 
that  this  pump  of  three  cylinders  is  to  be  driven  by  a  triple- 


326 


PUMPING   MACHINERY 


expansion  engine,  with  three  cylinders,  one  of  them  in  tandem 
with  each  of  the  water  cylinders.  The  plungers  are  single 
acting,  and  it  will  be  assumed  that  they  are  built  solid  to  give 
a  greater  pressure  on  the  downward  stroke. 

To  find  the  size  of  the  engine,  assume  that  the  steam  expands 
in  one  cylinder,  and  that  the  theoretical  card  is  of  the  form 
shown  in  Fig.  279.  This  card  assumes  no  clearance  and  com- 
pression, the  effect  of  these  together  being  equal  to  a  decrease 
in  area  of  about  3  per  cent.  If  the  initial  absolute  pressure 


FIG.  279.  —  Indicator  Cards  with  no  Clearance. 

be  called  pi,  the  pressure  at  the  end  of  expansion  px,  the  back 
pressure  p^  and  if  the  curve  of  expansion  be  taken  as  a  rectang- 
ular hyperbola,  the  mean  height  of  the  figure  becomes 


Mean  height  = 


This  quantity  is  multiplied  by  0.97  to  allow  for  the  effect 
of  clearance    and  compression.     There  is  another  effect  for 


DYNAMICS  OF  STEAM   END  327 

which  allowance  must  be  made.  This  is  the  effect  of  the  valve 
gearing  on  the  indicator  card.  The  corners  of  the  card  are 
rounded,  due  to  slow  valve  action,  and  there  is  some  change 
of  pressure  between  the  various  cards  as  shown  in  Fig.  280. 
This  means  that  a  diagram  factor  must  be  applied  to  give 
the  proportion  of  the  theoretical  card  which  may  be  actually 
obtained.  The  value  of  this  will  be  taken  95  per  cent  for  D- 
slide  valve  engines,  and  97  per  cent  for  Corliss  engines. 


FIG.  280.  —  Combined  Cards. 

The  probable  mean  effective  pressure  will  then  be  given  by 
M.E.P.  ^diagram  factor  X  .97  X  \px(i  +log?  |-lj  -pb  \  . 

The  size  of  the  low-pressure  cylinder  will  be  found  by  the 
formula 


33,000 

In  this  formula  the  values  of  P,  L  and  N  for  the  tandem- 
steam  cylinder  are  known  at  this  point,  and  A  may  be  found. 
If  the  engine  is  to  drive  by  gears  2LN  may  be  assumed,  next 


328  PUMPING  MACHINERY 

N,  after  which  L  and  A  are  computed.  In  some  cases  a  ratio 
of  L  to  D  is  assumed.  Having  the  size  of  the  low-pressure 
cylinder  the  size  of  the  other  cylinders  may  be  assumed  by 
taking  the  ratio  of  cylinder  volumes  from  practice.  A  better 
way  is  to  divide  the  area  of  Fig.  279  into  three  equal  parts, 
giving  equal  work  to  each  cylinder.  After  this  the  sizes  of  the 
cylinders  are  found  from  the  figure. 

If  the  receiver  is  assumed  to  be  very  large  the  pressure 
between  stages  will  be  constant. 

If  p\  is  the  initial  pressure  in  the  intermediate  cylinder, 
and  p"\  that  in  the  low-pressure  cylinder,  the  following  formula 
may  be  used  to  find  these,  assuming  equal  work: 


for  low  pressure; 


for  low  and  intermediate  together. 
From  these: 


log,  p"i  =i[log^,/>*2  +  2(     -i)] 


For 

/>i  =  150.3  lb-  gauge, 
px=  -6.7lb.  gauge, 
pb  =  -ii  7  lb.  gauge, 

the  following  pressures  result: 

p'i  =50.4  absolute; 
p"i  =  14.47  at  solute. 

These  values  have  been  marked  on  Fig.  279.  From  these, 
the  relative  dimension  of  the  cylinders  may  be  found,  since  the 
lines  ab,  cd,  and  ef  represent  respectively  the  volumes  of  the 


DYNAMICS  OF  STEAM  END  329 

high-pressure  cylinder,  of  the   intermediate-pressure  cylinder, 
and  of  the  low-pressure  cylinder,  hence 


These  ratios  are  the  theoretical  ratios  of  the  cylinders  if 
the  expansion  is  complete  in  each  cylinder.  If  there  is  free 
expansion  in  the  different  cylinders,  the  proportionate  drops 
in  each  are  assumed,  and  then  the  points  b  and  d  are  moved 

to  b'  and  d',  and  the  ratios  above  are  changed  to  —7  and  —7. 

With  such  ratios  the  volume  of  the  receiver  and  piping  is 
expressed  in  terms  of  the  cylinder  volumes  from  practice  as 
are  the  clearance  volumes.  So  soon  as  the  low-pressure  cylinder 
volume  is  computed  from  the  horse-power  equation,  the  volumes 
of  the  other  parts  are  all  known  and  the  combined  individual 
cards  may  be  drawn  as  in  Fig.  280.  The  receiver  and  piping 
between  each  cylinder  is  taken  as  about  250  per  cent  of  the 
cylinder  from  which  the  steam  is  discharging,  and  the  clearance 
on  the  three  cylinders,  to  be  5  per  cent  for  the  high-pressure 
cylinder,  2.8  per  cent  for  the  intermediate-pressure  cylinder, 
and  1.5  per  cent  for  the  low-pressure  cylinder.  These  values 
are  given  as  a  guide  so  that  these  quantities  may  be  computed, 
but  a  designer  after  making  a  number  of  designs  would  use 
the  tables  of  proportions  which  he  had  computed  from  previous 
experience. 

The  value  of  a  reheater  is  questioned,  a  matter  which  will 
be  taken  up  later. 

With  the  values  above,  the  data  for  the  problem  considered 
give  the  following  results: 

M.E.P.  =  .96  X  -9o[8(i  +loge  ip)  -3]; 
,=  25.2  Ibs.  per  sq.  in.; 

I.H.P.=46o  =  — 


33,000 


330  PUMPING  MACHINERY 

A  =2020  sq.in.; 
Z)=5i  in.; 
2LN  =2  Xfi  X6o  =  300  ft.  per.  min.,  which  is  not  too  high; 


8 


F,.p.    14.47    1.81' 

Fh.pi.8i       i 
Klp.      6.3  -3.48' 


The  strokes  are  all  the  same,  hence 
-  1  13.48:6.3, 


Since 

A  =51  in.; 
DA  =  20.33  in- 
A-  =37.8  in. 

The  first  approximation  of  cylinder  sizes  will  therefore  be 

20X30 
37X30 
51X30 

The  clearance  volumes  are 

470  cu.in. 

880      " 
890      " 

The  receiver  volumes  are 

23,530  cu.in. 
81,884      " 

From  these  volumes  and  pressures  the  theoretical  indicator 
cards  can  be  drawn  as  in  Fig.  280.  The  complete  expansion  of 
steam  to  the  receiver  pressure  is  found  in  most  pumping  engines, 
and  for  this  reason  these  cards  have  been  constructed  in  this 
manner.  The  cranks  are  assumed  at  120°,  and  the  pressures 
and  volumes  of  admission  to  the  various  cylinders  have  been 


DYNAMICS  OF  STEAM  END  331 

constructed  graphically.  The  receiver  volumes  and  clearance 
volumes  have  been  used  in  constructing  curves  where  possible; 
where  not  possible,  the  pressures  have  been  computed.  As  an 
example,  the  point  to  which  the  letter  a  (Fig.  280)  refers,  has 
been  computed  by  assuming  the  product  of  a  final  p  and  v 
as  equal  to  the  sum  of  the  original  products,  pv,  before  mixing, 

Pa(vah+VaI+VR)=pb(vbh+VR)+pcVcI. 

The  pressures  in  the  receivers  have  been  made  the  terminal 
pressure  of  expansion,  and  the  exhaust  has  been  arranged  to 
give  this  result. 

After  the  cards  have  been  drawn  they  have  been  used  to 
construct  cards  4  inches  long,  and  to  various  scales.  The 
cards  drawn  are  seen  to  be  of  different  size  from  the  original 
card  with  no  clearance;  the  H.P.  card  has  been  reduced,  the 
L.P.  card  increased,  and  there  is  little  change  in  the  I.P.  card. 
To  make  certain  that  the  steam  cylinders  are  of  sufficient  size 
the  I.  H.P.  of  the  total  engine  should  be  found  in  terms  of  the 
L.P.  cylinder  displacement,  assuming  that  the  volumes  of  the 
cylinders  are  of  the  ratios  given  originally.  Allowance  is  also 
made  for  the  piston  rod  by  considering  it  to  be  2j  per  cent 
of  the  piston  area.  After  finding  the  M.E.P.  from  each  theoretic 
card  in  Fig.  280  or  Fig.  281,  the  following  formula  may  be  used: 


J.H.P.  =        -  (pt+pi  M*  +P,)(Disp)low  N. 

33,000  Xi2\     6.3          6.3         Vv 

From  this  a  new  size  of  L.P.  may  be  computed. 

The  M.E.P.  from  the  indicator  cards  give  47  pounds  for 
the  high-pressure  cylinder,  15.9  pounds  for  the  intermediate, 
and  8.2  pounds  for  low-pressure  cylinder.  Substituting  these, 
the  equation  becomes 

i.Q75X6o  747      3.48  \ 

460=     v/3      -  (  |^-  +~-  15.0  +8.2  )D; 
33,000  X  12  V6.3      6.3  / 

Z)=  56,220  cu.in.; 

56,220 
Area  =  -^-=  1874  sq.m.; 

Diameter  =49  in. 


332  PUMPING   MACHINERY 

Hence 

Diameter  intermediate  =36.4  in.; 
Diameter  high  =  19. 5  in. 

Use  50,  37  and  20.  It  would  be  well  at  this  time  to  rede- 
sign the  cylinder  and  cards  of  Fig.  280  so  as  to  make  the  work 
on  each  cylinder  the  same. 

The  areas  of  these  cards  together'  will  be  greater  than  the 
area  of  the  water  cards  by  the  amount  of  friction.  The  fric- 
tion of  the  steam  and  water  pistons  and  the  friction  of  the 
stuffing  box  will  be  considered  to  be  three-fourths  of  the  total 
friction,  and  the  remaining  part  will  be  taken  uniformly  by  the 
.bearings  of  the  shaft. 

This  gives  22.8  H.P.  (460  —437.2)  as  the  total  to  be  absorbed, 
of  which  17.1  H.P.  is  used  in  the  packing  and  5.7  in  the  journals. 

17.1  H.P.  means  — -  —  =  9405  ft.lbs.  per  revolution. 

*  DO 

The  mean  pressure  from  this  over  2j-foot  stroke  would  be 

—  =627  pounds' for  each  unit  above  and  the  same  amount 
5X3 

below  the  zero  of  pressure,  as  this  force  operates  to  require 
work  and  always  to  hinder  motion.  This  quantity  is  to  be 
divided  by  the  areas  of  the  pistons  in  order  to  reduce  this  to 
the  same  basis  as  that  used  in  the  diagrams. 

The  weights  of  the  various  parts  will  be  assumed  to  be 
those  given  in  the  table  below.  The  weights  would  ordinarily 
be  determined  from  the  design  of  the  piston,  piston  rod,  cross- 
head,  and  connecting  rod,  which  would  be.  made  before  pro- 
ceeding with  this  part  of  the  work.  For  their,  design,  see 
Chapter  VIII. 

Weight  of  H.P.  piston 1 225  Ibs. 

Weight  of  I. P.  piston 900   " 

Weight  of  L.P.  piston 1400   " 

Weight  of  each  piston  rod 125    " 

Weight  of  each  cross-head 800   " 

Weight  of  each  connecting  rod 350   " 

Weight  of  4  rods  to  plunger  from  cross-head.  450   " 
Weight  of  plunger 5000   " 


DYNAMICS   OF  STEAM   END  333 

This  gives  as  the  sum  of  the  weights  of  the  reciprocating 
parts  of  each  cylinder  the  following: 

High^pressure  unit 6900  Ibs. 

Intermediate-pressure  unit 7575    " 

Low-pressure  unit 8075    " 

Reducing  these  to  pressure  per  square  inch  of  specific  steam 
cylinder  these  values  become: 

High-pressure  piston  22  Ibs.  per  sq.in. 
Intermediate-pressure  piston,  7  Ibs.  per  sq.in.: 
Low-pressure  piston,  4.1  Ibs.  per  sq.in. 
Friction  H.P.  piston  =  2.0  Ibs.  per  sqjn. 
Friction  I.P. 'piston  =0.575  Ib.  per  sq.in. 
Friction  L.P.  piston  =0.318  Ib.  per  sq.in. 

Taking  the  indicator  cards  from  the  separate  steam  and 
water  cylinders,  as  shown  in  Fig.  281,  it  is  necessary  to  add 
22  pounds  per  square  inch  for  the  high-pressure  unit  on  the 
down  stroke,  to  get  the  effective  pressure,  and  subtract  the 
same  on  the  up  stroke.  Moreover,  the  back  steam  pressure  on 
the  crank  end  must  be  subtracted  from  the  head  end  on  the 
down  stroke  in  order  to  get  the  net  pressure  from  the  steam. 
As  these  piston  areas  are  not  the  same  on  each  end  of  a  piston, 
to  reduce  them  to  the  same  scale  the  crank-end  pressures 

Ac 

should  be  multiplied  by  -p,  where  area  Ac  is  the  area  of  the 

A-h 

crank  end  of  the  cylinder  and  the  area  Ah  is  the  area  of  the 
head  end.  In  the  case  of  a  2o-inch  cylinder  with  a  3-inch  rod, 
this  ratio  is  o .  98,  and  hence  it  may  be  considered  as  unity. 
On  the  other  pistons  it  is  still  closer  to  unity. 

The  reciprocating  parts  require  accelerating  and  therefore 
a  force  equal  to 

w       w       /  I  \       • 

-a=-a>2r(cos  6+-  cos  20) 
g       g       \  n  9 

is  required  on  each  square  inch  of  piston  area,  if  w  represents 
the  weight  of  the  parts  per  square  inch  of  piston  area.  The 
values  of  a  for  the  different  positions  may  be  taken  from  Chapter 


PVMPINC  MACHINES? 


_Atmpspheric  Line 


. » «, 

(270  1240  J510 


Atmospheric  LineN 


NAtmosphcric  Line 


B. — Intermediate  Unit. 


FIG.  281. — Combined  Cards. 


DYNAMICS  OF  STEAM    END 


335 


V,  page  192,  and  plotted  on  stroke  position  after  multiplying 
by  -aj2r  for  each  cylinder.     These  curves   are  plotted  in  Fig. 

o 

281.    At  the  beginning  of  each  stroke  the  inertia  force  must  be 


^Atmospheric  Line 


C. — Low-pressure  Unit. 
FIG.  281. — Combined  Cards. 


subtracted  from  the  net  steam  force,  while  at  the  end  of  the 
stroke  it  is  added.  The  curve  repeats  itself  with  the  same 
numerical  values,  but  of  opposite  signs  for  210°  as  for  150°, 


336  PUMPING  MACHINERY 

for  240°  as  for  120°,  etc.  Hence  the  curve  need  not  be  redrawn, 
but  the  same  curve  may  represent  each  stroke,  provided  the 
sign  of  the  quantities  be  changed  for  the  two  strokes. 

Having  now  these  various  forces  which  act  on  the  system, 
it  is  a  simple  matter  to  find  out  how  much  of  it  remains  unbal- 
anced and  must  therefore  pass  from  this  system  into  the  others, 
or  into  the  fly  wheel  by  way  of  the  connecting  rod  and  shaft. 

Take  any  point  A  on  '  the  high-pressure  set  of  cards  and 
consider  this  to  be  on  the  up  stroke  (head  to  crank),  when  the 
fly  wheel  and  crank  are  assumed  to  be  above  the  steam  and 
water  cylinders  which  are  adjacent.  The  steam  pressure  on 
the  upper  side  is  aA,  on  the  lower  side  bA.  The  weight  is  cA^ 
the  inertia  is  dA,  the  friction  of  packing  is  Af  and  the  water 
pressure  is  hA  below  the  atmosphere.  The  water  pressure  is 
measured  from  the  atmosphere  because  the  air  pressure  is 
pressing  down  on  an  area  equal  to  the  area  of  the  plunger  minus 
the  area  of  the  steam  piston  rod  (a  small  quantity  compared 
with  the  piston  or  plunger  area)  and  this  balances  practically 
14.7  pounds  of  the  water  pressure.  The  pressure  on  the  water 
cylinder  has  been  reduced  to  a  figure  which  represents  pressure 
per  square  inch  of  steam  piston  area,  by  redrawing  the  diagram 

A 

of   Fig.    278,  using   for   pressures  P-j        —  ,  and   using  the 

•^  steam  end 

same  scale  as  used  for  the  other  diagrams.  The  net  pressure 
remaining  in  the  system  per  square  inch  of  steam  cylinder  area 
is 

P=aA  -bA  -cA  -dA  -Af-hA  =  CB. 

On  the  down  stroke  at  the  point  A  the  net  pressure  becomes 
Pl  =  a'A  -b'A  +  cA  +dA  -Af-h'A  =  CB'. 


(Note  that  arrows  showing  directions  -of  piston  movement 
aid  in  this  construction.)  The  figure  is  drawn  by  joining  points 
for  different  positions.  In  the  problem  assumed,  the  work 
of  each  steam  cylinder  was  taken  about  the  same  and  that 
of  each  water  cylinder  was  the  same,  hence  the  area  of  this 
resultant  figure  which  represents  the  net  work  done  by  the 


DYNAMICS  OF  STEAM   END 


337 


system  outside  of  itself,  must  just  equal  its  share  of  the  work 
done  in  journal  friction. 

Since  the  probable  cards  are  not  arranged  to  give  absolutely 
the  same  work  on  each  cylinder,  this  is  not  quite  true.  The 
diagrams  for  the  different  cylinders  have  been  made  of  the 
same  length,  but  to  make  them  of  proper  area,  the  high  pressure 
was  drawn  with  a  scale  of  50  pounds  per  square  inch  to  the  inch 


of  height,   the  intermediate  with 


3.48 


or  14.38  pounds  per 
50 


square  inch  to  the  inch  of  height  and  the.  low  pressure  with  ^ — 
or  7 . 95  pounds  per  square  inch  to  the  inch  of  height. 


FIG.  282. — Tangential  Effort  Construction. 

The  cards,  Fig.  281,  have  been  constructed  by  laying  off 
positive  pressures  above  the  base  line  EF  on  the  crank  head 
stroke,  while  for  the  head-crank  stroke,  positive  pressures  are 
laid  off  below.  This  gives  a  figure,  the  area  of  which  repiesents 
the  work  or  energy  transmitted  from  this  system  to  the  others 
through  the  connecting  rod  and  shaft. 

Having  now  the  force  per  square  inch  of  piston  which  is 
transmitted  to  the  crank  by  the  connecting  rod,  the  next  oper- 
ation is  to  find  the  value  of  the  force  produced  by  this  in  a 
direction  tangent  to  the  crank  arm.  In  Fig.  282  the  crank 
and  connecting  rod  are  shown  by  lines  in  a  given  position 


338 


PUMPING  MACHINERY 


of  the  crank.  The  end  A  of  the  rod  is  moving  horizontally  and 
therefore  it  is  moving  instantaneously  about  some  point  on 
a  perpendicular  to  the  path,  as  on  AC  for  the  position  shown. 
The  end  B  is  moving  in  a  "tangent  to  the  crank  circle  and  there- 
fore about  some  point  in  line  with  the  crank  radius  OBC.  Now 


30          60         QQ^   igft       i*m^>-re/T — '210 
H.P.  Effort. 


240  \   270        800 


330 


0        30 
120  H. 


60 
180  H. 


90  \120 


180      210        240 
300  H.  360  H. 

I.P.  Effort. 


120  H. 


FIG.  283. — Tangential  Efforts. 

one  end  of  the  rod  is  moving  about  a  point  in  one  of  the  lines 
and  the  other  end  about  a  point  on  another  line  and  the 
only  point  which  satisfies  both  of  these  conditions  is  the 
point  C.  That  is,  for  the  instant  considered  the  connecting 
rod  is  turning  about  the  point  C  as  much  as  if  it  were 
rigidly  connected  to  a  pivot  at  this  point.  If  then  it  turns 
about  this  point,  the  force  perpendicular  to  OB  produced 


DYNAMICS  OF  STEAM   END  339 

by  the  piston  force  P,  perpendicular  to  AC,  is  given  by  the 
equality 

TCB=PAC. 

The  effect  of  inertia  of  the  rod  is  such  that  this  equation 
is  not  strictly  true,  but  Jacobus  has  shown  in  a  paper  (A.S.M.E., 
XI,  pp.  492  and  1116),  that,  for  the  design  of  the  fly  wheel, 
the  approximate  method  of  considering  the  connecting  rod 
as  having  the  motion  of  the  piston  and  including  it  with  the 
other  reciprocating  parts  and  neglecting  the  real  inertia  effect 
at  this  point,  is  sufficiently  accurate. 

To  find  the  value  of  T  graphically,  lay  off  op  equal  to  the 
pressure  from  the  piston  rod,  and  draw  pt  parallel  to  the  connect- 
ing rod. 

The  triangles  otp  and  CAB  are  similar,  hence 


'  Ot  ~0t' 
=  OtBC] 

PAC=TBC~ 

hence  Ot  =  T. 

For  the  piston  pressure  P  at  the  cross-head  the  tangential 
effort  T  has  been  found.  This  is  done  for  several  points  in 
Fig.  283  and  plotted  on  a  line  representing  the  travel  of  the 
crank.  The  other  two  cards  are  drawn  as  shown. 

The  area  scale  of  these  figures,  Figs.  281  and  283,  is  equal 
to  the  product  of  the  two  linear  scales.  The  various  scales  are 
given  below: 

Card.       Length  Scale.                       Height  Scale.  Area  Scale. 

H.P.     f  ft.  per.  in.      50  Ibs.  per  sq.in.  H.P.  31.25    ft.-lbs.  .  per    sq.in. 

piston  per  in.  H.P.  area  per  sq.in. 

I.  P.     f  ft.  per  in.      14.38  Ibs.  per  sq.in.  of  I.  P.  8.99  ft.-lbs.  per  sq.in  of 

piston  per  in.  I.  P.  area  per  sq.in. 

L.P.     |  ft.  per  in.        7.95  lbs.persq.in.of  L.P.  4.96  ft.-lbs.  per  sq.in.  of 

piston  per  sq.in.  L.P.  area  per  sq.in. 

If  now  total  work  is  desired,  the  areas  of  the  diagrams  will 
be  multiplied  by  the  area  scale  and  the  area  of  each  piston. 


340  PUMPING  MACHINERY 

To  refer  these,  however,  to  the  L.  P.  area,  the   pressure  scale 

A 

for  the  H.P.  cylinder  will    be  multiplied  by  ~j?-     and    the 

**Lp 

A- 
I.  P.  scale  by  -—-.    When  this  is  done  it  is  found  that  all  area 

^I.P. 

scales  are  the  same. 


Since  then  the  tangential  diagrams  are  of  the  same  scale, 
they  may  be  combined  by  addition.  The  customary  arrange- 
ment of  cranks  as  seen  in  Fig.  312  is  to  have  them  at  120°  apart. 
Hence  the  three  positions  of  6=0°  will  be  placed  at  120°  apart 
as  in  Fig.  283  and  the  resultant  line  found.  The  area  of  this 
resultant  curve  will  be  the  energy  put  into  the  journal  friction 
and  if  the  height  .of  the  tangential  effort  required  for  this  be 
laid  off  above  the  zero,  it  will  be  found  that  this  will  be  the 
mean  height  of  the  curve  arid  the  amount  of  area  above  the 
line  will  be  equal  to  the  amount  of  area  beneath  the  line. 

These  areas  represent  work  or  energy  because  the  height 
represents  tangential  effort  or  force,  while  the  abscissae  represent 
crank  movement  or  motion  in  the  direction  of  the  force.  To 
reduce  these  areas  to  actual  total  force  they  are  multiplied 
by  the  area  scale  and  the  area  of  the  low-pressure  piston. 

Calling  the  areas  above  the  mean  line  excess  areas,  £lf 
£2,  ^3,  etc.,  and  those  below  the  line  deficiency  areas,  D^  D2, 
D3,  etc.,  the  variation  of  energy  in  the  fly  wheel  may  be  found 
as  follows: 

At  i  the  energy  in  the  fly  wheel  may  be  called  F;  then  by 
the  time  the  crank  has  rotated  to  2,  the  three  systems  have 
developed  an  excess  of  E\  units  of  energy  over  the  amount 
required  by  the  pump.  This  energy  must  be  absorbed  by  the  fly 
wheel  in  an  increase  of  speed.  By  the  time  3  is  reached,  D\  units 
have  been  abstracted,  and  so  on  for  the  various  points;  the 
energy  at  these  is  therefore  a^  follows; 


DYNAMICS  OF  STEAM  END  341 

Point.  Energy. 

1  F 

2  F+P^i 

3  F+El-Dl 

i  F+E1-Dl+E2-D2=F 

If  now  the  difference  between  the  greatest  and  the  least 
of  these  is  found,  that  quantity,  when  multiplied  by  the  area 
scale  and  the  area  of  the  low-pressure  cylinder  will  give  the 
amount  of  energy  to  be  stored  up  in  the  fly  wheel  to  change 
it  from  its  lowest  rate  of  speed  during  one  revolution  to  its 
highest  rate  during  that  revolution.  Calling  tl>is  area  difference 
AE  and  the  moment  of  inertia  of  the  fly  wheel  WR2,  where  R 
is  radius  of  gyration  in  feet  and  W  the  weight  of  the  wheel  and 
N'  and  N"  the  highest  and  lowest  rates  of  speed  in  R.  P.  M.  in 
one  turn  of  the  wheel,  the  following  results: 

AE  Xarea  scale  Xarea  L  P  piston 
«&* 


now  N'z  -  N"2  =  (N'  +  N" )  (Nf  -  N" )  =  2NAN 

where  N = mean  speed  =  R.  P.  M. 

AN  =  variation  in  speed.    . 

The  value  of  AN  depends  on  the  kind  of  machine  considered. 

N  N 

For  slow  pumps  —  may  be  used  while  —  would  be  better  for 

high  speed  pumps.     (In  the  case  of  spinning  mill  engines  and 

N 

electric  light  engines  the  value  —  is  used.) 

100 

Assuming  AN  and  the  R.P.M.  for  a  given  pump,  the 
quantity  WR2  is  given  by  the  formula.  If  R  is  known  for  a 
given  shape,  the  weight  W  could  be  found. 

In  the  case  of  the  ordinary  flywheel,  the  rim  is  the  most 


342  PUMPING  MACHINERY 

important  portion,  and  in  many  cases  the  rim  is  designed  to 
give  the  necessary  WR2.  In  the  case  of  the  flywheel  or  rim, 
of  width  b,  outer  radius  r\\  inner  rz\  with  hub,  of  width  blt 
outer  radius  r3,  inner  radius  r±  and  with  n  arms,  the  following 
results: 


WR2  for  rim  =  |  *  {w?xrbdr)r* 

Jn 


nbw 
-- 


2 

WP?  for  spoke,  uniform  elliptical  section, 


r2dr 

4 


4         3 
W- 


For  whole  wheel, 


In  this  equation  the  last  two  terms  may  be  found  by  taking 
data  from  practice,  and  then  by  assuming  r\  of  the  first  term, 
the  only  unknown  would  be  the  term  b  of  the  weight  W.  If 
desired  to  assume  b  this  may  be  done  and  then  the  unknown 
portion  of  the  first  term  would  be  r^  —  r24,  in  which  r2  is  assumed, 
fi4may  be  found. 


DYNAMICS  OF  STEAM  END  343 

If  two  flywheels  are  to  be  used,  \AE  is  used  for  each  wheel. 

The  first  tangential  diagram  of  Fig.  283  is  the  amount  of 
twist  which  is  transmitted  from  the  first  crank  to  the  first  fly- 
wheel, while  the  third  tangential  diagram  is  the  amount  which 
is  transmitted  from  the  third  crank  to  the  second  flywheel. 
The  middle  diagram  is  the  amount  which  is  transmitted  from 
the  center  crank  at  the  middle  to  the  two  wheels.  This  tan- 
gential effort,  when  multiplied  by  the  crank  radius,  gives  the 
twisting  moment  for  which  the  shaft  must  be  designed. 

The  net  piston  force  from  Fig.  281  gives  the  force  which  is 
applied  to  the  crank  shaft  to  produce  bending.  The  amount 
in  the  direction  of  the  piston  is  increased  by.  the  inclination  of 
the  connecting  rod.  The  maximum  increase  amounts  to  only 
i|-  per  cent,  and  so  this  increase  may  be  neglected  in  design, 
and  the  height  from  Fig.  281,  when  multiplied  by  the  pressure 
scale  and  the  area  of  the  piston,  will  give  the  force  which  causes 
bending  on  the  shaft.  However  this  may  be,  after  the  pump 
is  running,  if  anything  should  happen  to  the  machine  in  one 
of  the  inner  pump  barrels,  the  plunger  might  jam  and  then  the 
whole  steam  pressure  from  any  one  piston  would  reach  the  crank 
pin  and  hence  the  pin  should  be  designed  to  support  this. 


CHAPTER  VIII 
STEAM  END  DETAILS 

THE  steam  end  of  the  pump  has  been  illustrated  in  chapters 
III  and  IV  quite  fully,  and  it  remains  to  examine  peculiarities 
of  design.  The  action  of  the  steam  end  of  simplex  pumps  has 
been  fully  studied  and  an  explanation  was  made  of  the  action 
of  the  valve  of  the  duplex  pump. 

It  was  mentioned  that  in  order  to  operate  duplex  pumps 
properly  the  steam  valve  should  have  a  certain  amount  of  play  on 
the  stem.  This  is  accomplished  in  several  ways.  In  Fig.  284 


FIG.  284. — Steam  Valve  of  Duplex  Pump. 

the  method  of  using  a  central  nut  is  shown.  The  valve  is  made 
without  any  lap,  that  is,  the  valve  just  reaches  from  the  outer 
edge  of  one  steam  port  A  to  the  outer  edge  of  the  other,  which 
would  mean  that  as  soon  as  steam  was  cut  off  from  one  end 
the  other  would  open  for  steam;  reverse  the  pump  and  so  start 
the  other  pump,  which  in  turn  would  reverse  the  front  pump. 
By  having  the  valve  rod  slide  through  slots  in  the  projections 
on  the  back  of  the  valve  and  by  using  the  split  nut  of  proper 
length  to  drive  the  valve,  the  action  will  be  as  follows: 
Suppose  the  valve  is  moved  to  the  right  by  the  motion  of  the 
other  pump,  the  piston  controlled  by  this  valve  will  move  to 
the  right  and  it  will  start  to  move  the  valve  rod  on  the  other 

344 


STEAM   END  DETAILS 


345 


pump  to  one  side;  however,  the  valve  will  not  start  until  the 
piston,  which  is  moving,  reaches  a  point  near  the  end  of  its 
stroke,  whereupon  the  other  piston  starts  and  actuates  the 
valve  rod  of  Fig.  284  to  move  to  the  left.  The  valve  does  not 
operate  until  the  rod  has  traveled  the  distance  equal  to  the 
play  between  the  nuts  and  the  projection,  at  which  time  the 
valve  will  be  moved  to  the  left,  closing  the  left-hand  steam  port 
and  opening  the  right,  thus  starting  the  piston  on  its  return. 
The  time  taken  for  one  side  to  make  the  major  part  of  its  stroke 
is  the  time  taken  for  the  other  side  to  come  to  rest  at  the  end 
of  its  stroke  and  reverse.  In  this  way  there  is  a  very  steady 
discharge,  as  was  explained  earlier,  since  one  piston  is  moving 
at  full  speed  while  the  other  is  reversing.  There  is  a  small  period 
of  rest  at  the  end  of  the  stroke  when  the  cylinder  is  full  of  steam, 
while  the  other  piston  is  moving  to  throw  over  the  valve.  The 
driving  piece  C  should  be  made  in  two  parts  so  that  they  may 
be  jammed  to  hold  them  at  one  place  on  the  rod. 

Another  method  of  accomplishing  this  is  by  the  use  of  out- 
side striking  pieces,  as  shown 
in  Fig.  285.  This  arrangement 
is  better,  as  the  valve  may  be 
adjusted  at  either  end.  When 
this  adjustment  must  be  made 
often,  it  is  done  outside  of  the 
cylinder.  One  method  is  illus- 


FIG.  285.— Steam  Valve. 

trated  in  Fig.  286.     Here  the  valve  stem  A  is  provided  with 


FIG.  286.— Valve  Rod  Yoke. 

a  box  or  yoke  F  on  the  end.  This  is  driven  through  the 
cross  head  C  by  the  pin  B  attached  to  the  lever,  connected  to 
the  other  side  of  the  pump.  The  amount  of  play  is  regulated 


346 


PUMPING  MACHINERY 


by  the  set   screws  DD   which   are  fixed  in  place  by  the  jam 
nut£. 

In  Fig.  287  another  device  is  shown  to  accomplish  the  same 
result.  The  reverse  lever  A  is  driven  from  the  piston  rod  B  by 
a  link.  The  lever  is  pivoted  behind  the  valve  rod.  It  moves 


FIG.  287. — Driving  Mechanism  for  Valves. 

the  sleeve  C  by  means  of  a  link  and  this  sleeve  strikes  the  jam 
nuts  DD  and  moves  the  valve  rod  at  the  proper  time. 

These  last  two  devices  are  used  on  large  and  small  direct- 
acting  pumps.  They  are  used  at  times  on  the  rotating  steam 
valves  which  are  the  equivalent  of  D  valves. 


FIG.  288. — Triple  Expansion  Cylinders. 

The  cylinders  of  the  single  expansion  duplex  pumps  are 
quite  simple.  The  figures  of  chapters  III  and  IV  show  their 
construction.  The  action  of  the  five  ports,  the  form  of  valve 
and  piston,  the  arrangement  of  flanges  for  the  cylinder  heads, 
the  valve  chest  covers,  and  the  method  of  carrying  the  cylinders 
may  be  seen  by  reference  to  these  figures. 

The  arrangement  of  the  cylinders  of  a  triple  expansion  direct- 
acting  pump,  Fig.  288,  will  serve  to  illustrate  the  methods  of 


STEAM  END  DETAILS 


347 


building  these.  The  cylinders  are  cast  usually  with  single  walls, 
and  the  cylinder  ends  are  flanged  to  receive  the  cylinder  head. 
Heads  may  be  cast  solid  with  the  walls.  This  latter  method 
serves  to  cut  down  the  number  of  joints  to  be  packed  and  does 
not  materially  increase  the  cost  of  construction.  The  heads  of 
the  cylinders  are  dished  or  cored  to  receive  the  nuts  at  the  end 
of  the  piston. 

The  arrangement  of  piston  rods  shown  in  the  figure  is  designed 
to  reduce  the  number  of  inside  packing  boxes  and  make  it 
possible  to  examine  each  piston  without  disturbing  any  cylinder. 

When    the    cylinders    are   jacketed,    several   methods    are 


"a 

u 

(i 

8 

FIG.  289. — Leavitt  Cylinder  and  Jacket. 

available.  The  jacket  may  be  made  in  the  cylinder  casting 
or  it  may  be  made  by  the  use  of  a  liner. 

When  the  jacket  is  made  in  one  piece  with  the  cylinder 
barrel,  a  casting  difficult  to  produce  results,  the  outer  shell 
is  apt  to  cool  first  and  when  the  inner  part  cools,  strains  are 
set  up  which  crack  the  casting.  To  prevent  this  Mr.  E.  D. 
Leavitt,  Jr.,  devised  the  scheme  of  casting  the  outer  shell  of 
the  jacket  with  a  division  in  the  outer  wall,  Fig.  289.  The 
opening  thus  made  was  closed  by  a  copper  ring  A,  with  one 
corrugation.  This  ring  was  attached  to  the  edges  of  the  opening, 
by  tap  bolts  with  a  sharp  ring  on  the  lower  face  of  the  head, 
which  was  driven  into  the  copper  and  made  a  steam-tight  joint. 
The  edge  of  the  copper  was  caulked  against  the  cast  iron. 

The  Snow  Steam  Pump  Company  makes  jackets  at  times 
with  a  cast  corrugation  around  the  center  of  the  shell  to 
allow  for  this  expansion  and  contraction.  This  .cylinder, 


STEAM  END  DETAILS 


349 


Fig.  290,  is  jacketed  on  the  heads  as  well  as  on  the  barrel. 
The  steam  enters  the  barrel  jacket  at  A  and  the  condensation 
is  removed  at  B.  The  condensation  from  the  heads  is  removed 
by  the  drains  C,  which  in  some  cases  serve  as  admission  pipes 
also,  the  system  being  the  equivalent  to  a  single  pipe  system 
of  heating.  The  air  valves  at  DD  relieve  the  jacket  of  air. 


FIG.  291. — Valve  Gear  with  Cut-off  Controlled  by  Speed  Governor. 

The  method  of  attaching  the  heads  by  stud  bolts  is  clear.  The 
stuffing  box  is  shown  at  K. 

The  half  cross  sections  through  the  valves  and  center,  Fig. 
290,  show  how  the  steam  entering  through  EE  passes  the 
Corliss  steam  valves  to  the  exhaust  passages  FF,  which  are 
cross  connected  to  the  manifold  G. 

The  ribs  at  H  and  /  are  to  stiffen  the  casting  where  passages 
are  made. 

The  steam  and  exhaust  valves  and  gear  are  shown  in  this 
figure  and  in  Figs.  291,  292.  In  these  the  knock-off  cam  L 
and  arm  M  are  seen.  In  Fig.  291  the  position  of  L  is  fixed  by 


350 


PUMPING   MACHINERY 


the  position  of  the  governor  through  the  rod  N  from  the  governor, 
and  depends  on  the  speed,  while  in  Fig.  292  the  position  is 
controlled  by  the  pressure  in  the  force  main  from  the  pump. 

The  steam  pipe  0  connects  the  inlets  E  leading  to  the  valve 
chests,  while  the  exhaust  pipe  G  connects  the  exhaust  outlets. 
The  dash  pots  QQ  are  carried  by  the  exhaust  piping. 

The  jacket  is  sometimes  constructed  by  introducing  a 
liner  within  the  cylinder,  Fig.  293.  In  such  design  the  liner 


FIG.  292. — Valve  Gear  with  Cut-off  Controlled  by  Pressure  Regulator. 

may  be  made  of  a  hard  iron,  while  the  main  casting  is  of  a 
soft  cast  iron,  which  is  less  liable  to  crack.  The  liner  is  fastened 
at  its  lower  end  by  bolts,  while  at  its  upper  end  a  copper  ring 
is  used,  fastened  by  tap  bolts  with  sharp  circular  lips  on  the 
lower  faces.  The  jacket  is  supported  by  rings  on  its  outer 
surface  in  contact  with  rings  projecting  from  the  inner  surface 
of  the  shell.  The  surfaces  of  the  projections  are  the  only  portions 
of  the  inner  bore  of  the  cylinder  or  the  outer  surface  of  the 
liner  which  require  turning.  Moreover,  by  making  these  of 


STEAM   END  DETAILS 


351 


decreasing  diameter  toward  the  closed  end  of  the  cylinder 
and  by  making  each  surface  slightly  tapered  the  liner  may  be 
introduced  with  little  or  no  forcing  and  yet  will  be  held 
tight. 

The  cylinder  shown  in  Fig.  291  is  intended  for  a  horizontal 


FIG.  293. — Independent  Liner. 


FIG.  294. — Head  Valves. 


engine,  but  the  same  construction  would  be  used  if  for  a  vertical 
one.  To  cut  down  the  clearance  by  eliminating  the  long  passages 
extending  from  the  straight  valve  face  to  the  circular  bore  of 
the  cylinder,  engines  have  been  built  with  the  valves  in  the 
heads.  Fig.  294  shows  such  an  arrangement  for  Corliss  valves. 
Many  companies  restrict  this  head  valve  to  the  low-pressure 
cylinders  only.  The  system  has  the  great  advantage  of  simplify- 


352 


PUMPING   MACHINERY 


ing  the  cylinder  casting,  but  for  small  high-pressure  cylinders 
this  is  not  so  important. 

The  simple  form  of  cored  piston,  Fig.  295,  with  two  piston 
rings  is  a  very  common  type.  It  is  attached  to  the  tapered 
end  of  the  piston  rod  by  a  nut.  A  collar  is  used  to  keep  the 
rod  in  position  and  support  part  of  the  thrust. 

The  half  rings  shown  beside  the  piston  illustrate  the  two 
forms  of  rings  used  with  small  pistons,  one  of  constant  cross 
section  and  the  other  with  variable  cross  section. 

Fig.  296  is  the  type  used  for  large  cylinders.     The  main 


m 

FIG.  295. — Steam  Piston. 


portion  A  is  attached  to  the  piston  rod  as  shown,  and  from 
the  ring  B  a  number  of  bolts  radiate,  supporting  a  bull  ring  C. 
This  bull  ring  contains  a  groove  in  which  is  placed  a  ring  made 
of  a  series  of  overlapping  sections.  These  are  held  out  against 
the  cylinder  bore  by  helical  springs  in  the  cavities  D.  The 
follower  plate  E  is  held  in  position  by  the  tap  bolts  FF.  The 
ring  B  is  supported  by  a  series  of  webs  from  the  central  hub  of 
the  piston. 

Fig.  297  illustrates  another  one  of  these  pistons  in  which 
the  sections  are  held  out  by  carriage  springs  and  the  bull  ring 
has  two  additional  grooves  which  are  filled  with  a  bearing 
compound,  so  that  the  weight  of  the  piston  may  be  taken  on 


STEAM  END  DETAILS 


353 


a  substance  which  will  not  wear  away  the  bore,  but  all  wear  will 
occur  in  the  ring  which  may  be  turned  as  soon  as  the  cast  iron 
of  the  bull  ring  comes  in  contact  with  the  bore.  In  this  way 


these  bearing  rings  will  last  for  some  time,  as  it  will  take  eight 
or  ten  movements  to  use  up  the  complete  circumference. 
In  Fig.  298  the  bull  ring  is  provided  with  two  sectional 
rings. 


354 


PUMPING   MACHINERY 


These  figures  with  those  of  the  previous  chapter  illustrate 
the  method  of  attaching  pistons. 

The  simple  stuffing  boxes  have  been  discussed  under  the 


FIG.  297. — Piston  with  Sectional  Ring. 

head  of  water-end  details.  The  metallic  packing  is  quite  com- 
mon and  as  an  example  of  one  form  Fig.  299  is  given.  In  this 
the  rings  A  are  forced  into  the  conical  cups  BE  by  the  springs 


FIG.  298. — Piston. 

C  through  the  ring  D.  The  cup  B  has  a  ground  steam-tight  joint 
against  the  ball-ended  ring  E.  The  ball  E  fits  the  socket  F. 
The  socket  F  can  not  be  moved,  the  ball  and  socket  allow  for 
the  alignment  of  the  rod  and  the  ground  joint  between  B  and 


STEAM  END   DETAILS 


355 


E  allows  the  rod  to  move  across  the  cylinder  when  wear  of 
the  piston  or  cross  head  makes  this  necessary.    This  is  possible 


FIG.  299. — Metallic  Packing. 

since  the  rings  or  cups  E  and  F  are  all  of  a  larger  inside  diameter 
than  that  of  the  rod. 

DESIGN  OF  CYLINDER 

The  thickness  of  the  cylinder  is  given  by  the  same  formula 
as  suggested  in  chapter  VI.     (See  Figs.  290,. 293  and  294.) 


5000 


t\v  =^iiv  =o .  6  thickness  of  single  plate  cover. 

Thickness  of  single  plate  cover  =  o.  55^  \— , 

/vi  =i"  to  i"  for  liner  jackets. 
=  i"  to  3"  for  cast  jackets. 

The  valve  chest  walls  and  covers  have  their  thicknesses 
determined  by  the  cylinder  thickness  formula  or  the  flat  sur- 
face formula. 

/v«  =0.35^ \- (or  0.55^  \-  for  round  plates), 
s  \  s  /. 


356  PUMPING  MACHINERY 

Flanges  here  are  made  i  .  2$tvii  in  thickness. 

The  diameter  of  the  bolt  should  vary  with  the  size  of  cylinders. 
|"  bolts  are  the  smallest  to  be  used  and  these  may  be  increased 
to  ij"  on  very  large  cylinders,  f-"  and  i"  bolts  are  common 
on  cylinders  of  20".  After  the  diameter  is  assumed  the  number 
of  bolts  is  found  by  the  formula 


#  =  area  at  root  of  thread  of  bolt  in  sq.in.  from  table  of 

standard  threads. 
St  =  allowable  stress  in  pounds  per  sq.in. 

=4000  Ibs.  per  sq.in.  to  allow  for  straining. 
n  =  number  of  bolts. 

p  =  maximum  steam  pressure  in  Ibs.  per  sq.in. 
d  =  diameter  of  cylinder  in  inches. 
Flange  width  3^  at  least. 

Pitch  of  bolts  <  40  \- 

STEAM  PASSAGES 
The  area  of  steam  passages  is  given  by 

2LNA 

6000  ' 

while  the  exhaust  passages  are  given  by 

2LAN 

a  =  --  . 
4500 

In  these  a  is  the  area  of  cross  section  of  tne  passage  in  square 
inches;  L,  the  stroke  in  feet;  N,  the  revolutions  per  minute,  and 
Ay  the  area  of  the  piston  in  square  inches. 

The  length  of  the  port  is  made  about  o  .  8  diameter  of  cylinder, 

then,  width  =  —  5-3.     The  value  6000  for  the  velocity  of  steam 
o.oa 

should  only  be  used  to  find  the   steam  passages.     The  port- 
opening  is  found  by  using  this  value  when  cut-off  is  at  0.7 


STEAM  END   DETAILS  357 

stroke.  For  0.25  stroke,  the  port-opening  is  found  by  using 
18,000  for  the  velocity.  For  cut-offs  between  these  use  a  pro- 
portional amount. 

CLEARANCE 

The  clearance  distance  is  used  to  allow  for  wear  in  the  connect- 
ing rod  and  bearings,  unevenness  in  the  castings  of  the  piston 
and  head,  and  errors  in  alignment.  The  clearance  distance 
is  made  from  ^"  to  ^"  in  different  sizes  of  engines,  the  first 
being  used  for  strokes  of  15"  and  under.  The  latter  for  strokes 
to  6  feet. 

The  clearance  volume  is  given  by 


LA 

• 

In  these  LI  and  LI  are  the  length  of  the  steam  and  exhaust 
passages. 

The  clearance  has  an  important  effect,  as  has  been  shown 
in  many  engine  tests,  so  that  it  is  desired  to  keep  the  value 
of  this  as  small  as  possible.  The  following  table  will  give  some 
idea  of  the  variation  of  this  quantity. 

QiYY-iTVIa  TT-nrrlnck'  Clearance  Volume  in  Percent- 

Simple  Engine .  age  of  Cylinder  Displacement. 

D-Slide  valve 6  to  15 

4  valve  engines 2  to    4 

Corliss  valves 2  to    4 

Piston  valves 7  to  15 

Compound:  H.P.  L.P. 

Slide  valve 7  to  15        5  to  10 

Corliss  valves  in  side 2  to    4    1.3  to   3 

Triple:                                                       H.P.  I.P.             L.P. 

Slide  valves 10-15  5-10          5-7 

Corliss  valves 1.5-1.75  1.25-2.0  1.5-2.5 

Corliss  valves  in  head. .                     4— i          ^-i 


358  PUMPING  MACHINERY 

PISTON  PROPORTIONS 

(See  Figs.  295  and  296.) 

The  design  of  the  piston  is  empirical  and  the  following 
proportions  are  recommended  by  various  authors: 


Flat  disc  i=  --  -  for  cast  iron. 

100  ... 


f 

for  steel. 


150 

This  value  will  be  used  as  the  unit  for  all  pistons;  and  dimen- 
sions on  the  figures  are  in  terms  of  this: 

B  =0.40  and  for  large  cylinders  ^/LD  (inches). 
Thickness  of  faces  =£'=0.5^ 
Thickness  of  face  when  follower  ring  is  used  =o.& 
Thickness  of  center  boss  -/"  =  0.5*  +*'". 
Thickness  of  ribs  =  *'"  =  0.3* 
Distance  from  edge  to  ring,  b'  =Q.$t. 
Thickness  of  junk  ring  or  follower  plate  /iv=o.5^. 
Diameter  of  junk  ring  bolts  ^'  =  0.25  t. 
Number  of  webs  =o.iD  +  2. 
Thickness  of  bull  ring  at  edge  tv  =  o.6t. 
Thickness  of  bull  ring  under  packing  ring  ffi  =  .$t. 
Thickness  of  spring  packing  rings  tvii=o.o^D. 
Width  of  spring  packing  rings  &"  =  £"  to  f". 
Thickness  of  solid  piston  wall  beneath  ring  /Vlii  =t. 
Diameter  of  piston  ring  =  1.0157). 

The  basis  for  the  proportions  used  for  the  piston  rings  above 
as  given  by  Unwin  in  his  "Machine  Design,"  is  as  follows: 

Assume  the  piston  ring,  Fig.  295,  to  be  loaded  with  p'  pounds 
pressure  per  square  inch  and  to  be  changed  by  this  from  a 
radius  R  at  any  point  to  the  radius  r  of  the  cylinder.  If  the 
width  of  the  ring  is  6"  the  resultant  pressure  causing  bending 
at  the  angle  of  6  from  the  end  after  fitting  into  the  cylinder  is 


=  2r  sn 


STEAM   END   DETAILS  359 

and  its  moment  is 


=  \2rsin  ^ 


SI 

This  bending  moment  is  equal  to  -~  or 

* 


. 

6  2 

The  change  in  curvature  also  depends  on  this  moment  or 

I_£_M 

r~R~EI' 

This  last  expression  is  used  to  determine  the  variation  in 
curvature  when  I  is  constant,  or  the  variation  of  /  if  the  curvature 
before  springing  into  the  cylinder  is  constant.  The  first  equation 
is  used  to  determine  the  thickness  at  the  point  in  the  ring  half 
way  from  the  ends.  In  this  case  #  =  180°  a  ad 


. 
6 

Using  6000  for  5  and  2  pounds  for  p'  the  following  results: 


6 


. 
oooo 

If  now  the  section  is  made  uniform,  as  on  the  left  of   Fig. 
295,  the  value  of  R  will  have  to  be  different  at  various  points. 


0 

2-b  b  r2  sin2  — 
i     £  2 

R~^~E  Xo.ooooi8b"~r*' 


360  PUMPING  MACHINERY 

for  />'  =  2lbs.    and    £  =  15,000,000. 

6 

Ar2  sin2  — 
I      I  2 


2  —  Ar2  sin2  —     270  —  4  sin2  — 

For  different  values  of  6  the  values  of  R  are  given  below: 

0  =  o°  30°        60°         90°        120°        150°       180° 

#  =  1.000?'     1.003^     1.004;'     1.007^     i.our     1.014?     i.oi5r. 

The  curve  of  the  original  curvatures  is  drawn  in  Fig.  295; 
the  radius  for  180°  being  used  for  a  length  .of  arc  equal  to  15° 
of  the  cylinder  circumference  on  each  side  of  the  180°  position. 
This  gives  the  point  a,  and  on  the  radius  to  this  point  a  new 
center  is  taken  equal  to  that  for  #  =  150°.  This  arc  is  made 
equal  to  30°  of  the"  cylinder  circumference,  giving  the  point  b. 
On  the  radius  to  -this  point  a  new  center  is  taken  for  the  curve 
to  c,  then  in  a  similar  manner  d,  e,  f  and  g  are  found.  When 
this  ring  is  sprung  together  until  the  points  touch,  it  will  just 
fit  in  the  cylinder  and  will  exert  a  uniform  pressure  of  2  pounds 
per  square  inch. 

If  R  is  constant  the  expression  for  curvature  becomes 


or 


STEAM  END   DETAILS  361 

Hence  the  following  results  in  terms  of  t]^' 


6=0 

5 

10 

15 

30 

45 

60 

tvii=o 

.124 

.197 

•257 

.405 

•529 

.630 

90 

120 

150 

180 

•794 

.908' 

.971 

i 

The  thickness  t]£Q  is  made  the  same  in  this  as  in  the  former 
method. 

PISTON  ROD 

The  piston  rod  is  designed  as  described  in  chapter  VI, 
page  314. 

For  the  remainder  of  this  chapter  the  maximum  pressure 
on  the  piston  will  be  called  P,  where 


A-P- 

**t—  c  ' 


A  i  =  area  at  root  of  thread; 
A  =area  of  main  rod. 

Shoulder  $"  to  J"  with  taper  of  3"  in  12". 

STUFFING  Box 

The  stuffing-box  design  has  been  given  in  chapter  VI, 
page  201. 

CROSS  HEADS  AND  CONNECTING  RODS 

The  cross  head  is  necessary  to  guide  the  piston  rod  and  keep 
it  from  deflecting  under  compression  and  also  to  take  the  side 
thrust  from  the  connecting  rod.  The  amount  of  this  side  thrust 


362 


PUMPING   MACHINERY 


is  shown  by  a  force  diagram  in  Fig.  300.  The  three  forces,  P 
from  the  piston  rod,  P'  from  the  connecting  rod  and  7?  from 
the  guides  are  in  equilibrium.  R  and  P'  will  be  the  largest 

for  a  given  value  of  P  when  a  is  a  maximum.    Sin  a  =— .    ac  is 

ac 

constant.  Hence  sin  a  is  a  maximum  when  ab  is  a  maximum. 
This  occurs  when  ab=*oa  or  0  is  90°.  In  this  case  the  length  of 


FIG.  300. — Force  Diagram. 

P'  in  force  polygon  is  n  times  the  length  of  R  and  the  following 
relations  hold: 


P'  = 


=P, 


(n=  ratio  of  connecting  rod  length  to  crank  arm.) 
Although  P  rarely  continues  its  full  value  until  the  crank 
has  moved  90°,  since  cut  off  occurs  earlier  than  this,  it  is  well  in 
design  to  consider  this  to  happen  and  design  accordingly. 
This  load  might  be  developed  if  the  full  pressure  were  con- 
tinued. 

The  form  of  cross-head  used  in  pumps  when  the  water  end  is 
in  tandem  with  the  head  end  of  the  steam  cylinder  or  where  the 
pump  is  driven  from  another  part  of  the  shaft,  as  in  a  triplex 
pump,  is  shown  in  Fig.  301. 


STEAM    END  DETAILS 


363 


i      ^ 
i 

| 

M 

364 


PUMPING  MACHINERY 


This  type  is  not  the  principal  type  used,  but  it  is  of  good 
form  and  is  therefore  given. 

The  main  casting  A  is  made  like  a  rectangular  box  with  one 
end  open  and  bosses  BB  placed  on  two  sides,  while  a  longer  boss 
or  tube  C  is  placed  on  the  side  opposite  the  open  end.  On  the 
upper  and  lower  sides  of  the  main  casting  A  are  placed  slippers 
or  shoes  EE  which  are  supported  by  wedges  D.  The  lower 
face  of  the  wedge  D  rests  on  the  horizontal  surface  of  A ,  but  its 
upper  face  is  inclined  and  supports  the  inclined  face  of  E.  The 
wedge  may  be  moved  by  turning  the  bolt  F,  which  is  j  ammed 
after  moving  the  wedge  far  enough  to  lift  the  slipper  out  from 


FIG.  302. — Cross-head. 

the  casting  an  amount  equal  to  the  wear.  The  slippers  are 
faced  with  a  bearing  metal  such  as  babbit. 

The  pin  G  is  slightly  tapered  where  it  passes  through  the 
sides  of  A.  This  makes  it  possible  to  have  the  pin  tight  without 
driving.  The  pin  is  fastened  by  a  nut  to  the  plate  H  which  is 
held  in  position  by  two  tap  bolts  //  while  the  set  screws  // 
jam  the  plate  so  that  there  is  no  danger  of  the  bolts  //  backing 
off. 

The  piston  rod  is  screwed  into  the  boss  C  and  a  jam  nut 
keeps  it  from  backing  off. 

The  cross-head,  Fig.  302,  illustrates  a  form  which  is  used 
when  four  guide  bars  are  employed.  The  main  casting  consists 
of  two  blocks  A  A  connected  by  the  yoke  B.  The  cross-head 


STEAM  END  DETAILS 


365 


pin  C  is  separate  from  the  casting  and  is  held  in  position  in  the 
grooves  made  for  it  by  the  pin  D.  The  piston  rod  fits  into  the 
tapered  hole  at  E  and  is  held  in  place  by  a  cotter. 


Plunges 


FIG.  303. — Crank  and  Reciprocating  Parts. 


FIG.  304. — Two-rod  Cross-head. 

Fig.  303  shows  the  arrangement  of  the  steam  piston  and 
water  plunger  when  these  are  in  tandem  with  the  crank-shaft 
between  them.  When  this  is  the  case  the  cross-head  has  to  be 
so  arranged  that  the  rods  which  are  used  to  cross  over  the 


.  :  FIG.  305. — Pump  End  Cross-head. 

crank  and  shaft  may  be  fastened  to  the  cross-head.  As  shown, 
this  is  accomplished  by  two  rods,  one  above  and  behind  the 
center  line  of  the  piston  rod  and  the  other  below  and  in  front 
so  fixed  that  they  are  symmetrical  about  the  center  line. 

The  cross-head  at  the  steam  end  is  showm  in  Fig.  304.     It  is 


366 


PUMPING  MACHINERY 


STEAM  END  DETAILS 


367 


368 


PUMPING  MACHINERY 


of  the  same  form  as  Fig.  301  with  the  addition  of  the  arms  A  A. 
The  water  end  cross-head,  Fig.  305,  is  only  used  to  prevent 
buckling  of  the  rods  when  under  compression  and  hence  is  not 
so  large  as  the  other. 

In  some  cases  the  designer  prefers  to  use  four  rods  "sym- 
metrically placed,  and  in  such  a  case  the  form  used  is  that 
shown  in  Fig.  306.  This  form  is  modified  on  account  of  the 
great  distance  between  the  frames  in  which  the  guide  surfaces 


FIG.  309. — Strap  End. 

must  be  placed.  The  sliding  blocks,  as  seen  in  the  figure,  are 
cast  blocks  fitted  to  the  pin  at  the  center,  which  passes  through 
the  main  portion  and  forms  the  wrist  pin.  These  blocks  are 
used  for  guiding  only,  the  largest  part  of  the  steam  forces 
going  directly  from  the  piston  rod  to  the  pump  rods. 

The  types  of  connecting  rods  used  on  these  machines  are 
shown  in  Figs.  307  and  308.  In  Fig.  307  the  main  forging  is 
made  with  boxes  at  each  end  and  is  circular  in  section  at  A, 
being  turned  from  A  to  B  as  a  cone  with  its  faces  slabbed  off. 
The  box  ends  C  and  D  are  fitted  with  brasses  EE  which  are 


STEAM  END  DETAILS 


369 


tightened  up  for  wear  by  the  wedges  FF.  The  wedges  are 
moved  by  the  bolts  G  which  are  jammed  by  the  nut  on  their 
opposite  faces.  The  wedges  are  arranged  to  shorten  one  end, 
as  wear  occurs  while  the  other  end  is  lengthened. 

In  Fig.  308  the  rod  is  made  with  a  conical  section  in  which 
the  slabbing  of  the  sides  occurs  near  the  crank  end.     In  this 


FIG.  310. — Marine  End. 

rod  the  box  at  the  crank  end  is  made  by  inserting  a  block  G 
between  two  fork  ends  HH  and  bolting  it.  This  is  necessary 
when  the  rod  is  used  with  the  center  crank  A  of  Fig.  312.  The 
rod  307  could  only  be  used  with  the  overhung  crank  B  of  Figs. 
311  and  312. 

The  strap  end  309  and  the  marine  end  310  may  be  used  with 
center  cranks.  In  the  strap  end,  Fig.  309,  the  strap  A  is  held 
on  the  square  end  of  the  rod  by  the  gib  B  and  the  key  or  cotter 


370 


PUMPING  MACHINERY 


C  inserted  in  a  slot  through  the  end  of  the  rod.  The  key  is 
prevented  from  coming  out  of  the  slot  by  a  set  screw,  and  the 
end  of  this  set  screw  engages  with  the  key  in  a  groove  so  that 
the  burr  formed  does  not  interfere  with  the  removal  or  adjust- 
ment of  the  key.  The  brasses  and  other  parts  are  the  same  as 
in  the  previous  rods.  The  marine  end,  Fig.  310,  has  two  brasses 


c 


FIG.  311.— Crank  Shaft. 


A  A  held  to  the  club  end  B  of  the  rod  by  a  keeper  C  and  two 
bolts.  The  liners  D  allow  the  bolts  to  be  made  tight  and  jammed 
without  gripping  the  crank  pin.  When  wear  occurs  these  are 
reduced  in  thickness. 

The  crank-shaft  for  a  two  cylinder  engine  is  shown  in  Fig. 
311,  while  in  Fig.  312  a  three-crank  shaft  is  illustrated.  In  these 
the  arms  BB  may  be  driven  on  and  keyed  or  in  some  cases  they 


FIG.  312. — Three-crank  Pin  Shaft  with  Return  Cranks. 

are  forged  solid  in  one  piece.  The  bearing  journals  are  at  DD 
and  at  times  the  shaft  is  enlarged  at  the  point  of  attachment  of 
the  flywheel,  as  at  C.  A  key-way  is  left  for  the  flywheel.  These 
shafts  are  sometimes  forged  hollow  to  reduce  weight  and  to 
remove  metal  which  is  of  little  value. 

DESIGN  OF  CROSS-HEAD,  CONNECTING  ROD  AND  SHAFT 

The  designs  of  the  cross-head,  connecting  rod  and  shaft  are 
interdependent  so  that  it  is  well  to  take  up  the  designs  together. 


STEAM   END   DETAILS  371 


THE  CROSS-HEAD 

The  pressure  on  the  cross-head  to  be  carried  from  the  piston 
is  P  pounds,  P'  pounds  from  the  connecting  rod  and  R  from 
the  guides,  where 


If  the  connection  with  the  piston  rod  is  made  by  a  thread, 
the  amount  of  thread  to  be  covered  by  the  boss  of  the  cross- 
head  should  be  equal  to  the  diameter  of  the  rod.  If  a  key  is 
used  to  key  the  rod  to  the  cross-head  the  thickness  of  the  key 
(t,  Fig.  302)  should  be  o.zd  and  the  width  is  given  by 


2Xo.2dS8' 

The  pin  should  be  investigated  for  crushing  and  the  rod 
for  bearing  by  the  formulae: 

o.2dXdSr>P 


The  cross-head  pin  may  be  designed  empirically  and  then 
investigated  for  bearing,  bending,  and  deflection.  A  good  rule 
is  to  make  the  length  of  the  pin  (/  —  Figs.  301  and  302)  equal  to 

x/P7                                                                        x/P7 
,  and  the  diameter  (df  —  Fig.  301  and  302) . 

P' 

The  bearing  pressure  jj,  then  becomes   1200  pounds  per 
let 

square  inch,  which  is  allowable  for  wrist  pins. 


372  PUMPING    MACHINERY 

The  investigations  for  stress  and  deflection  are  given  by  the 
formulae: 

s-    / 

*~>8  — 


nd'2' 
2 

4 


100  _     ^TTtf4' 

384^ 

These  are  usually  fulfilled,  but  it  is  well  for  the  student  to 
investigate  expirical  design. 

The  area  of  the  shoe  is  designed  so  as  to  give  a  pressure  of  40 
pounds  per  square  inch  in  fast-running  engines  with  cast  iron 
slippers  and  250  pounds  in  slow  engines  when  the  shoe  is  lined 
with  babbit  metal  and  80  pounds  in  case  of  cast-iron  slippers 
against  cast-iron  guides  for  slow  running. 

A'=l'w'=j,. 

p'  is  the  allowable  pressure  for  any  given  problem. 

The  relative  values  of  /'  and  w'  are  fixed  by  the  design. 

\lr  =wf  may  be  used. 

The  pin  is  held  by  the  sides  of  the  main  casting.  In  Fig.  301 
the  supporting  surface  is  zd't". 

P' 


The  details  of  the  main  casting  (Fig.  301 )  are  given  in  terms 
oid'. 

Thickness  side,      t'"  =  0.3^',  but  must  be  not  less  than  J", 
Thickness  boss,      *"  =  *•'" + \", 
Thickness  bottom,  ?v  =0.25  d', 
Thickness  front,      F  =  o.2$d'. 


STEAM  END   DETAILS  373 

In  the  cross-head  shown  in  Fig.  302  the  design  is  somewhat 
different.    Here  the  thickness  must  be  calculated. 


™ 

~ 

<*"'  = 


The  cross-head  pin  is  cut  off  for  20°  at  top  and  bottom  as 
this  is  not  of  much  value  and  it  distributes  the  oil  better 
when  so  treated. 

The  thickness  of  the  side  blocks  tvii  is  assumed  and  then 
investigated  for  strength.  The  piece  is  under  combined  bending 
and  direct  stress  and  is  investigated  as  follows: 

Fig.  302  shows  the  general  arrangement  of  the  forces  and 
their  positions.  The  thickness  tv{i  may  be  assumed  and  then 
the  center  of  gravity  c,  moment  of  inertia  /,  and  area  of  section 
A',  shown  at  side,  are  found.  The  point  of  application  of  the 

P 

load  —  is  at  tyvl  from  the  edge  so  that  there  is  a  lever  arm  of 

c~^tv{  causing  bending.    On  the  crank-head  stroke  the  maxi- 
mum tension  is 

P      Pc-±F{c 


*~ 


2l 


on  the  other  stroke  the  maximum  compression  is  the  same. 
Since  the  tensile  strength  is  usually  less  than  that  of  the  com- 
pressive  strength  for  the  material  used  for  cross-heads,  the 
investigation  for  St  is  the  only  one  made.  If  S*  is  found  to  be 
greater  than  the  allowable  St,  the  section  must  be  increased  or 
cast-steel  used  in  place  of  cast-iron. 

The  yoke  is  designed  as  a  cantilever  beam.    At  the  center 
it  is  best  to  assume  dviii  and  compute  I"'  by  the  formula 

p 

- 


374  PUMPING  MACHINERY 

For   a  section  at  x   from  the  center,  the   thickness  div  is 
assumed  and  the  distance  /iv  is  found  by  the  formula: 


In  Fig.  305  the  arms  are  designed  as  cantilever  beams  by 
assuming  t  and  finding  w  from  the  formula: 


In  Fig.  306  the  bending  moment  is  %Pb.  Although  the  force 
P  is  in  general  not  the  whole  force  on  the  steam  piston,  yet  the 
design  should  be  made  to  carry  this  load  since  that  force  might 
come  upon  it. 

The  length  of  the  crank  pin  is  determined  by  the  heating 
caused  by  friction.  The  total  piston  pressure  P  will  produce 
the  amount  of  friction  //P  where  /*  is  the  coefficient  of  friction. 
This  force  moves  through  the  distance  rdd.  The  work  per 
minute  will  be 

W  = 

If  this  is  divided  by  the  product  of  the  length  and  the  diameter 
of  the  pin  I'd',  the  result  will  be  the  amount  of  work  per  square 
inch  of  projected  area  of  pin.  This  quantity  then  becomes 


. 

f2* 

If  for    I     PdO  the  approximate  value,  2nPmean  is  used,  the 
J° 

above  expression  becomes 


W 


I' 


By  using  the  values  of  «,  N,  P  and  /  in  a  number  of  engines 
which  have  run  without  heating,  a  mean  value  of  w  may  be 
found  and  from  this  value  on  substitution  a  formula  for  /  may 
be  derived: 


since  Pmean  = 


STEAM   END   DETAILS  375 

I.H.P.X33,ooo 


2LN   .      ' 


These  two  formulae  may  be  used  to  determine  /'  by  making 
7^  =  0.00003  and  K'  =o.j  to  i.o. 

The  value  of  u.  is  0.04  for  proper  lubrication  and  o.io  for 
poor  lubrication. 

After  the  length  of  the  crank  pin  is  determined  its  diameter 
is  found  to  care  for  strength,  bearing  pressure  and  deflection. 
These  three  are  considered  separately.  If  the  crank  pin  is 
overhung,  the  pin  acts  as  a  cantilever  beam.  Then 


Pi' 


32 


Forbearing 


In  this  p  is  the  unit  bearing  pressure.  For  slow  engines 
Unwin  allows  800  to  900  pounds  per  square  inch.  While  on 
fast  engines  the  value  is  from  500  to  800,  in  marine  engines 
400  to  ^oo.  For  deflection  the  formula  is 


i_ 
*  El  ~ 


The  amount  of  deflection  is  limited  to  o.oi  inch.  If  the  crank 
pin  is  a  center  pin,  the  design  for  strength  is  to  be  considered 
with  the  design  of  the  shaft,  and  for  that  reason  the  shaft  design 
will  now  be  studied. 

In  beginning  to  design  a  shaft,  assumptions  must  be  made 


376 


PUMPING  MACHINERY 


of  the  various  lengths.  The  length  of  the  pin  /'  has  been  found 
and  this  will  serve  as  a  guide  for  certain  other  parts.  It  may 
be  assumed  that  crank  arms  have  the  width  /'  and  that  the 
bearings  have  a  length  2,1' .  With  these  dimensions  in  view  an 
approximate  sketch  is  made.  Fig.  313  shows  this  for  an 


f*^f  •»*-£•* 

E- 


X 


FIG.  313.— Shaft  Sketch. 

overhung  crank  and  Fig.  314  for  a  triple  crank-shaft,  both  of 
approximate  form. 

The  distances  A  are  assumed  of  sufficient  size  for  the  fly- 
wheel or  to  bring  the  cylinders  of  the  multicylinder  engine  at 
the  proper  distances  apart.  The  quantities,  a,  b,  c,  d,  e,  etc.,  are 
now  found  in  actual  lengths  in  inches. 

From  the  diagrams  of  Fig.  281  the  greatest  steam  pressure 


^-+  —  A  —  ^ 

—e- 

«_^  A  —  1_,_ 

xi             bxc 

XT                    IX 

i          d          ,i. 

—^u  f  ^  

FIG.  314.— Shaft  Sketch. 

can  be  found  and  the  crank  position  at  which  it  acts  can  be 
constructed,  giving  a  diagram  as  Fig.  315  from  which  P'  is 
found.  This  force  acts  on  the  crank  pin;  but  its  equivalent 
Pf  acts  at  the  center  of  the  shaft  in  connection  with  a  couple. 
This  is  seen  in  the  figure  by  adding  two  equal  and  opposite 
forces  at  the  center  of  the  shaft.  One  of  these  combines  with 
P'  to  form  a  couple  which  produces  twisting  only  while  the 


STEAM  END   DETAILS 


377 


other  force  left  at  the  center  of  the  shaft  produces  bending 
only.  The  other  piston  forces  on  a  triple  crank-shaft  may  be 
found  by  obtaining  the  positions  of  the  other  pistons  from  a 
diagram  similar  to  Fig.  282,  and  then  from  the  piston  position 
obtaining  the  pressures  from  Fig.  281  by  multiplying  the  heights 


p' 


FIG.  315. — Forces  on  Shaft. 

by  the  spring  scale  and  the  area  of  the  piston.  The  direction 
and  magnitude  of  the  rod  pressure  are  found  in  a  manner  similar 
to  that  shown  in  Fig.  315. 

The  weights  of  the  flywheels  are  now  taken  from  the  deter- 
mination of  chapter  VII,  and  after  resolving  the  connecting 


FIG.  316. — Loading  Diagram. 

rod  pressures  into  vertical  and  horizontal  components,  the 
shaft  is  treated  as  a  beam,  first  subject  to  vertical  forces  and 
then  to  horizontal  forces.  After  the  bending  moments  or  bend- 
ing moment  diagrams  for  these  are  computed,  the  resultant 
moments  may  be  found  and  the  maximum  determined.  The 
loading  diagrams  for  a  vertical,  overhung  engine  is  given  in 
Fig.  316.  In  these  figures  the  forces  are  arranged  as  if  they 


378 


PUMPING  MACHINERY 


were  acting  in  the  same  direction.  They  may  in  reality  be 
reversed  and  when  this  is  the  case  the  sign  of  the  term  involving 
any  such  should  be  changed. 

The  critical  points  in  Fig.  313  are  at  the  left  support  and 
under  the  load. 

These  may  be  computed 


Mv  (at  left  support  )=Pv'a, 
Mh=Ph'a. 


At  load: 


The  resultant  of  Mv  and  Mh  =  VMV2  +Mh2. 


FIG.  317. — Loading  Diagram. 

The  largest  of  these  is  the  one  which  fixes  the  size  of  the 
shaft. 

In  the  case  of  the  three-cylinder  engine,  the  shaft  is  a  contin- 
uous beam  of  three  spans  loaded  with  concentrated  loads  and 
hence  the  theorem  of  three  moments  must  be  applied  to  find 
the  moment  at  the  supports. 

The  general  form  of  the  theorem  for  this  case  is 


STEAM  END  DETAILS  379 

The  equations  for  this -particular  case  for  which  the  loading 
diagram  is  Fig.  317,  are  as  follows: 


From  these  four  equations  the  values  of  M  over  each  support 
may  be  found  and  then  the  shear  to  the  right  of  each  support 
is  given  by 

I/      ^2,-Mj      W 
d          K2J 

M3v-M2     P'v 

V2=     ~7~~+T' 

MA.—  Mo..       W 


f  '    2- 

(These  are  of  this  form  since  load  is  at  center.*) 
The  moments  at  the  load  will  then  be 


V2e 


This  same  operation  is  performed  for  the  horizontal  forces, 
remembering  that  the  terms  involving  W  are  zero.  From  the 
values  of  the  positive  and  negative  moments  the  resultants  are 
found  and  the  maximum  determined.  In  this  manner  deter- 
mination should  be  made  for  various  positions  of  the  crank  and 
the  maximum  of  the  maxima,  ascertained. 

This  may  be  the  correct  method,  but  a  better  method  would 

*The  general  form  for  Vi  is      Vi  =  Mz~Ml+—  +P(i-Ki). 


380  PUMPING  MACHINERY 

be  to  assume  the  engine  fixed  on  the  dead  point  on  one  cylinder 
with  full  steam  pressure,  and  have  the  full  net  steam  pressure 
on  one  of  the  others  the  same  while  the  third  has  exhaust  pressure. 
With  this,  work  out  the  equation  above.  A  determination 
should  also  be  made  of  full  boiler  steam  on  the  first  crank  alone 
with  no  pressure  on  the  other  pins.  In  these  problems  the 
pressures  from  the  pistons  might  be  considered  as  acting  directly 
on  the  pins  with  connecting  rods  of  n  =6,  as  the  effect  of  angu- 
larity is  not  great  enough  to  affect  the  result. 

After  making  a  number  of  investigations  the  values  of  M 
at  the  various  pins  are  known. 

From  the  tangential  effort  diagram,  Fig.  283,  the  twisting 
moment  of  any  shaft  can  be  found  as  well  as  the  twist  at  the 
various  pins.  Twisting  moment  is  equal  to  the  tangential 
effort  multiplied  by  crank  radius.  The  diagram  of  H.P.  cylinder 
gives  the  twist  from  the  first  crank  to  the  first  flywheel,  and 
the  third  diagram  from  the  L.P.  cylinder  gives  the  twist  going 
to  the  flywheel  near  it.  The  amount  going  to  the  flywheels 
is  given  by  the  combined  diagram,  and  if  one  half  of  this 
is  subtracted  from  the  H.P.  diagram  the  amount  which 
remains  shows  the  quantity  going  to  the  second  crank  from 
H.P.  The  twist  from  the  L.P.  diagram  may  be  treated  in  the 
same  manner  and  the  amount  received  by  the  second  crank 
from  L.P.  is  given.  The  quantities  from  the  two  ends  when 
added  should  give  the  amount  from  the  I.  P.  diagram.  If  the 
amount  from  the  H.P.  diagram  reaching  the  second  crank  is  of 
the  same  sign  as  that  on  the  I.  P.  diagram,  the  whole  twist  from 
H.P.  must  be  transmitted  through  the  second  crank  pin.  If 
the  amount  of  twist  from  the  I.  P.  diagram  is  less  than  that  from 
the  H.P.  but  of  different  sign,  then  the  arithmetic  difference 
is  transmitted.  In  this  way  a  diagram  may  be  drawn  for  the 
twist  which  is  transmitted  across  the  center  crank. 

If  now  from  these  diagrams  the  bending  moments  and  twist- 
ing moments  are  obtained  at  various  points  in  the  revolution 
for  the  pins  and  shafts  below  the  wheel,  the  values  of  T\  = 
could  be  found. 


It  has  been  shown  by  Grashoff  that  when  bending  and 


STEAM  END  DETAILS  381 

torsion  occur  at  the  same  time  in  a  shaft,  the  shaft  should  be 
designed  as  subject  to  a  bending  moment  of  MI  where 


when  M=  bending  moment, 

T  =  twisting  moment. 

MI  =  combined  equivalent  bending  moment 

Rankin  develops  a  slightly  different  formula  by  considering 
maximum  stress,  in  place  of  maximum  deformation  and  arrives 
at  the  formula 


Mr.  J.  J.  Guest,  within  the  last  few  years,  has  examined  bodies 
subject  to  tension  and  torsion  and  finds  that  the  controlling 
force  is  not  one  of  tension  but  of  shear,  and  so  he  recommends 
that  the  combined  equivalent  shear  be  used  in  designing  when 
these  two  stresses  occur.  This  would  give  for  the  moment 
formula: 


The  diameter  of  the  shaft  or  pin  may  now  be  found  by  this 
formula. 

It  is  to  be  noted  in  passing  that  when  a  triple-expansion 
three-cylinder  engine  drives  a  pump  or  other  machine  at  one 
end  of  the  shaft,  the  crank  pin  nearest  the  point  of  driving 
transmits  the  sum  of  the  twists  from  the  cylinders  placed  in 
front  of  it. 

The  investigation  for  bearing  should  be  made  with  center 
cranks  at  this  point,  although  d'  is  so  large  here  on  account  of  the 
bending  moment  that  it  will  rarely  be  found  necessary  to  enlarge 
d'  to  that  given  by  the  formula: 

-4 

Having  now  the  diameter  of  the  crank  pin  for  the  various 
forms  of  cranks,  with  the  length,  the  boxes  for  the  ends  may  be 
designed. 


382  PUMPING   MACHINERY 

The  proportions  recommended  by  Unwin  are  given  in  Fig. 
318,  where  the  unit  is 


when  the  pin  is  an  overhung  pin,  but  with  center  pins,  the  unit  is 


The  connecting  rod  proper  is  subject  to  stress  produced  by 
four  causes:  ist,  Direct  tension  from  the  piston  pull  when  the 
piston  moves  toward  the  head;  2nd,  Direct  compression  on  the 
return;  3d,  The  bending  stresses  produced  by  the  column 
flexure,  and  4th,  The  bending  stresses  produced  by  the  whipping 
of  the  rod  as  it  changes  the  direction  of  motion  at  the  top  and 


FIG.  318. — Box  for  Connecting  Rod. 

bottom  positions  of  the  crank  in  a  horizontal  engine.  The  stress 
in  the  first  case  is  equal  to  that  in  the  second,  and  as  the  second 
and  third  are  included  in  the  ordinary  column  formula,  investi- 
gation by  that  formula  is  sufficient.  The  fourth  stresses  are 
produced  by  the  inertia  of  the  rod,  and  if  m  be  the  mass  for  any 
unit  length  of  the  rod,  and  a  the.  maximum  acceleration,  ma 
represents  the  inertia  force  over  that  length  tending  to  bend 
the  rod.  The  acceleration  in  the  rod  is  a  maximum  near  the 
crank  position  90°  from  the  dead  point.  The  rod  is  pivoted  at 
the  cross-head  pin,  and  hence  the  acceleration  varies  as  the 
distance  from  that  point.  If  the  connecting  rod  is  uniform  in 
section,  the  effect  of  inertia  is  the  same  as  if  there  were  a  beam 
loaded  with  a  distributed  load  which  increased  from  zero  at  one 


STEAM   END   DETAILS 


383 


end  to  a  certain  value  at  the  other,  as  in  the  upper  part  of  Fig. 
319.  The  bending  moment  is  a  maximum  at  O.577/,  and  it  is 
at  this  point  the  dangerous  section  for  bending  occurs. 

If  the  section  of  the  rod  is  not  uniform,  but  increases  toward 
the  crank  end,  the  loading  does  not  vary  along  a  straight  line 
but  along  some  curve,  as  shown  in  the  lower  part  of  Fig.  319. 
In  this  case  the  dangerous  section  for  the  bending  is  further  out. 

Now  the  dangerous  section  of  the  column  is  at  J/,  but  owing 
to  the  bending  effect  from  inertia,  consider  o.6/  as  the  dangerous 
section  with  a  rod  of  varying  cross-section  for  both  bending, 


FIG.  319. — Inertia  Load  on  Connecting  Rod. 

due  to  the  column  and  to  inertia.     If  the  rod  is  uniform  the 
dangerous  section  would  be  between  \l  and  0.577^. 

The  stress  produced  at  the  dangerous  section,  treating  the 
rod  as  a  column,  is  given  by  the  formula: 


while  the  stress  by  the  inertia  bending  is  given  by 

Me 


Now  this  latter  could  only  be  determined  after  knowing 
the  shape  of  the  rod,  and  hence  the  shape  would  have1  to  be 
assumed  for  a  first  approximation.  In  place  of  this  such  a 
value  of  S  in  the  column  formula  may  be  taken  that  this  inertia 
effect  would  be  covered.  If  such  a  value  is  used  and  the  cross-sec- 
tion deterrhined  from  formula,  the  shape  is  found  at  the  danger- 
ous section,  which  is  considered  to  be  at  o.6l. 


384  PUMPING  MACHINERY 

The  rod  may  now  be  designed.  The  cross-head  end  of  most 
rods  is  made  smaller  than  the  remainder  of  the  rod  and  is  designed 
to  take  direct  tension  or  compression.  This  is  therefore  designed 
first.  If  the  rod  is  to  be  uniform  in  section  this  end  is  not  de- 
signed, as  the  size  is  determined  by  the  cross-section  at  the  center. 

The  force  which  this  rod  may  have  to  stand  is,  as  was  men- 
tioned under  crank  pins, 

-       * 


Vn2-i 

In  many  engines  the  cut  off  occurs  before  the  crank  has 
reached  its  mid  position,  but  it  is  best  to  use  this  value  as  the 
valve  may  be  reset  so  that  the  pressure  could  be  carried  to  the 
90°  crank  position. 

The  area  at  this  smallest  section  is 

Pf 
4000' 

The  value  5=4000  being  found,  from  practice,  to  be  a  fair 
average. 

If  the  rod  is  rectangular, 

A"  =  b"d", 
in  which  V  or  d"  is  assumed; 

..        n—=A",         -     ':.     ,;:- ,-...;, V/ 

gives  the  diameter  if  the  neck  is  circular.  The  connecting  rod, 
considered  as  a  column  in  the  plane  of  oscilla- 
tion, has  round  ends,  while  in  a  perpendicular 
plane  it  has  fixed  ends.  In- investigating  rods 
of  standard  engines,  considering  the  dangerous 
section  of  rods  similar  to  Fig.  320  as  a  rectangle, 
which  would  circumscribe  the  section  and  not 
as  the  actual  section,  it  was  found  that  a 

FIG.  320. — Standard  r  ,  ,    .         . 

Rod  Cross-section.    stress   of  4ooo  pounds   used  in   the  column 
formula  would  give  sections  found  in  modern 
engines.     Hence  this  includes  the  effect  of  whipping  as  well  as 
the  column  effect. 


STEAM  END  DETAILS  385 

Taking  now  the  plane  of  oscillation,  for  the  section  at  o.6/, 
the  following  results  : 

P'  4000 

'AJTf=          4      &' 


25000  r2 

or  in  case  the  width  is  the  same  as  the  diameter  at  the  smallest 
section, 

Pf  4000 

^7r/=     ZTTJL' 

+  25000  d'"* 

12 

From  these  d'"  at  the  dangerous  section  is  found. 

Now  in  the  plane  perpendicular  to  the  plane  of  oscillation, 
consider  the  rod  as  a  column  with  fixed  ends.  This  is  not  strictly 
true,  as  there  is  some  play,  but  since  in  this  plane  there  is  no 
whipping  effect,  the  value,  5=4000,  will  equalize  the  decrease 
of  strength  due  to  the  rod  not  having  strictly  fixed  ends  in  that 
direction. 

P'  4000 


d"d'" 


25000  ^2 

12 


Solving  this  for  d'",  another  value  is  found  and  the  larger  of 
these  two  is  taken. 

Barr  treats  rectangular  sections  another  way.  The  danger- 
ous section  is  at  the  center,  and  if  df"  is  made  2d",  columns.  of 
equal  strength  in  the  two  planes  result,  hence  only  solve  for 
d"f  once.  The  formula  used  is  that  of  Euler,  which  reduces  to 

d"=c\/Di9 

For  steel:    0=0.057  mean, 

=0.07    maximum, 
=0.043  minimum, 


386  PUMPING  MACHINERY 

In  order  to  overcome  the  whipping  effect  df"  is  made  greater 
than  zd".    The  mean  value  is 


—  4.^"  maximum, 
=  2.2d"  minimum. 

The  factor  of  safety  here  is  13.  In  this  case  the  taper  is 
obtained  by  giving  the  rod  at  the  neck  the  same  cross-section 
as  the  piston  rod  and  tapering  from  the  point. 

Theoretically  the  taper  need  not  extend  beyond  a  point 
at  o.6l,  but  in  high-speed  engines  there  is  some  chance  for  the 
crank  pin  to  seize  and  then  the  connecting  rod  becomes  a  canti- 
lever beam  loaded  at  the  end  which  might  snap  off.  Hence  in 
this  class  the  rod  taper  is  usually  carried  out  to  the  crank  end. 
In  slow-speed  engines  the  rod  very  often  tapers  each  way  from 
the  center  and  is  circular  in  cross-section. 

The  brasses  or  boxes  EE,  Fig.  307,  are  next  considered  in  the 
light  of  the  above  rod  design  and  an  endeavor  is  made  to  make 
the  distance  I"  of  Fig.  318  equal  to  d".  This  result  would  make 
the  forging  of  the  rod  simpler,  and  if  it  can  be  done  by  making 
the  thickness  of  the  flange  slightly  greater  or  less  by  any 
amount  than  that  shown  in  Fig.  318,  these  should  be  changed 
to  accomplish  this  end. 

CONNECTING   ROD   ENDS 

The  connecting  rod  end  is  next  considered.  The  sides  of  the 
box  end  carry  tension;  the  end,  bending;  the  rod  end,  compress- 
ion. The  bolts  of  the  marine  end  carry  tension;  the  keeper, 
bending,  and  the  club  end,  compression.  The  strap  end  has  to 
be  designed  for  tension  and  bending  in  the  strap,  compression 
in  the  rod  end,  and  shear  and  compression  in  the  gib  and 
key. 

Box  end.  (Figs.  307  —  308.)  Assume  the  diameter  of  the 
bolt  attached  to  the  wedge  £"  or  £"=^iv. 


STEAM   END  DETAILS  387 

Find  b  from  brass  and  if  possible  make  it  equal  to  the  thick- 
ness of  the  rod  d".     Then 


The  rod  end  could  be  designed  as  a  beam  from  the  formula 


6 

Since  the  bending  of  this  rod  end  would  send  the  load  to  the 
comers  it  may  be  well  to  neglect  the  design  of  this  as  a  beam, 
determining  only  the  thickness  at  the  corners  for  shear  and 
making  r  =  i^vi  and  using  a  circular  arc  for  the  ends,  as 
shown,  for  appearance, 


The  wedge  should  be  sufficiently  high  for  crushing  and  also 
about  three- fourths  the  height  of  the  opening  of  the  rod  in  length 
so  that  the  brass  will  be  well  supported. 


The  bolt  at  the  end  of  rod  shown  in  Fig.  308  supports  all  of 
the  load,  hence 

^p7 

-dv. 


**  r : 

This  should  be  designed  for  crushing  and  the  fork  end  should 
carry  tension 

P' 


P' 


2S,(6-^)' 

This  may  require  a  thickening  because  of  the  large  bolt  used . 
Such  a  result  is  shown  in  Fig.  308. 


388  PUMPING  MACHINERY 

p' 
Marine  end.     Bolt  area  at  root  of  thread,  =— ; 5- 


4 
Breadth  keeper,  bl  =di  =  amount  from  box  design. 

/i  =  d'  +  2t  +  di  (t  from  brass  ). 

C     Z,      /     9 

Thickness  keeper,  |P'/i  =    '  *  1  . 

Length  of  keeper,  l,+2di  =/„. 
Crushing,  P'  <SJ//b1. 
Strap  end.     Key  and  gib  design: 

'"1'"  =  2S/V"  =  25,  (ft  -  6"'X'  =P'. 


c 

srf-*|  (in  general) 


Taper  of  key  f "  per  foot. 

The  key  and  gib  are  each  made  £/'"  in  width  at  top,  when  key 
is  first  entered. 

In  the  rod  end  the  height  t  is  usually  much  greater  than  2t,  and 
hence  there  is  no  need  of  investigating  for  crushing  between  the 
key  and  the  rod  end.  If  this  were  not  so  an  investigation  would 
be  made. 

The  end  of  the  strap  is  designed  in  the  same  manner  as  the 
end  of  the  box. 


STEAM  END   DETAILS  389 

BEARINGS 

The  bearings  of  the  shaft  are  designed  for  sufficient  bearing 
area  to  care  for  the  load.  Using  the  vertical  shears  at  each  side 
of  the  bearing  in  the  continuous  beam  design  the  algebraic 
difference'  of  these  would  be  the  reaction,  and  if  this  is  found 
in  the  horizontal  and  vertical  planes  the  resultant  R  is  given 

by 


As  an  example  in  Fig.  317,  the  shears  at  the  left  of  the 
supports  have  been  found  as  Vi,  V2,  V$y  etc.  On  the  right 
of  these  supports  the  shears  will  be 

Pv'  at  right  of  first. 

Vl  at  left  of  first. 

Fi  +  TF  at  right  of  second. 

F2  at  left  of  second 

V2+PV"  at  right  of  second,  etc. 

The  first  reaction  from  the  vertical  forces  will  be  P'  —  V\. 
The  various  .reactions  from  the  vertical  forces  will  be  found  in 
the  same  manner;  in  like  manner  the  horizontal  reactions  and 
from  them  the  resultant. 

In  Fig.  316  the  reactions  are: 

Right  horizontal,  —PhT- 
Right  vertical,  -?A 
Left  horizontal,  Ph     ,     . 
Left  vertical,  Pvf^~ 


Right  resultant,  =R  =  ^Ph'2  ~  +  (  -Pv' % 


T  TJ  1   i  *•-»•  '    *       *»-».       *     •'•'        I       v      \  i      -r-k      *     ^       1       ^  *    -r-r-r    \ 

Left  resultant, 


390 


PUMPING  MACHINERY 


Having  the  pressure  allowable,  the  diameter  of  the  shaft 
at  the  journal  (from  stress  design)  and  the  load  on  the  bearing, 
the  length  is  given  by 

R_ 

~  pd"' 

For  p  Unwin    recommends  from  150  to  450,  the   larger 


FIG.  321. — Main  Bearing. 

values  being  used  with  slow-running  engines,  in  which  there 
may  be  a  reversal  of  load  as  near  the  stroke  end.  For  fast  engines 
with  shafts  carrying  a  steady  load  in  one  direction  the  low 


STEAM  END    DETAILS 


391 


value  should  be  used.  200  pounds  is  an  average  value  which 
may  be  used. 

The  bearings  of  large  pumps  are  shown  in  Figs.  321  and 
322.  They  are  usually  made  in  four  parts: — one  bottom 
brass  A,  two  cheek  pieces  B,  and  one  top  piece  C.  This  is 
done  so  that  wear  may  be  taken  up  horizontally  and  vertically. 

In  building  the  boxes  for  the  main  bearing  it  is  well  to 
make  the  lower  section  circular  at  the  bottom,  so  that  when 
necessary  to  renew  this  section  it  may  be  taken  out  by  blocking 


\ 

o 

FIG.  322. — Bearing  and  Frame. 

up  the  shaft  and  twisting  the  box  around.  Of  course  such  a 
section  does  not  resist  the  effect  of  friction  to  turn  the 'brass, 
nor  can  a  shim  or  liner  be  introduced  beneath  the  circular  bottom 
as  it  could  be  with  the  square  base.  If  the  box  can  be  moved 
along  the  axis  of  the  shaft  by  the  removal  of  bolts  or  caps  this 
would  enable  one  to  remove  the  box  easily,  and  in  this  case  a 
flat  base  could  be  used,  as  in  Fig.  322. 

The  proportional  parts  of  these  boxes  and  bearings  are 
shown  in  Figs.  321  and  322.  The  boxes  are  to  be  grooved  so 
that  oil  will  be  distributed  and  in  many  cases  it  will  pay  to 
deliver  oil  under  pressure  to  the  main  bearings.  In  any  case 


392  PUMPING  MACHINERY 

a  steady  oil  stream  should  be  used,  the  oil  being  filtered  after 
using. 

The  thickness  and  projection  of  flanges  on  bearing  brasses 
are  made  J/. 

If  desired  the  surface  of  the  central  portion  of  the  box  which 
supports  the  box  may  be  cut  away  throwing  the  load  to  the 
ends  of  the  box.  When  this  is  done  the  width  of  the  removed 
portion  is  J/. 

Where  collars  or  crank  bosses  wear  against  the  flange  of  the 
box  a  slight  amount  of  play  is  given  to  the  shaft  by  leaving  a 
small  amount  of  clearance. 

The  bolts  holding  the  top  cap  should  be  strong  enough  to 
hold  the  cap  in  place,  and  also  in  the  case  of  horizontal  engines, 
to  carry  part  of  the  thrust  of  the  piston  across  the  top  of  the 
opening  in  the  frame.  This  may  be  done  by  projections  on  the 
side  of  the  top  cap,  as  shown  in  Fig.  322,  when  the  bolts  are 
only  used  to  keep  the  cap  in  place.  The  exact  amount  of  this 
force  required  to  hold  the  cap  in  place  cannot  be  computed,  so 
that  this  part  of  the  design  will  be  empirical,  and  the  following 
number  of  bolts  will  be  used. 

Shaft  Diameter.  No.  of  Cap  Bolts.       Diameter. 
6"  2  ij-'* 

8  2  I| 

12  4  If 

18  4  If 

24  42 

These  bolts  are  often  made  T-headed  and  fitted  into  pockets 
in  the  sides  of  the  spaces  left  to  receive  the  boxes. 

The  wedge  bolts  for  lifting  are  usually  two  in  number  with 
one  or  two  set  screws  in  addition,  so  that  the  bolts  may  be  used 
to  raise  the  wedge  while  the  set  screws  are  used  to  force  it  down 
after  the  bolts  are  loosened,  or  they  may  be  used  to  hold  the 
cheek  piece  in  place  by  tightening  them  up  against  the  pressure 
from  the  bolts.  These  bolts  and  set  screws  may  be  made  f" 
in  diameter  for  boxes  of  shafts  up  to  12"  and  i"  beyond  that. 


STEAM  END  DETAILS  393 

FLY  WHEEL 

The  fly  wheel  is  usually  made  as  shown  in  Figs.  323  and  324. 
That  shown  in  Fig.  323  is  in  two  sections  while  that  of  Fig.  324 
is  in  five  sections. 

The  two-section  wheel  is  bolted  together  at  the  shaft,  and 
at  the  rim  it  is  held  by  a  link  A  which  passes  around  projections 
BB  on  each  end  of  the  sections.  There  are  two,  three  or  four  of 
these  rings  used  at  each  joint.  These  projections  are  usually 


FIG.  323.— Two-part  Fly  Wheel. 

made  by  forming  a  ring  shaped  cavity  in  the  fape  of  the  rim 
so  that  the  ring  is  flush  with  the  surface. 

In  Fig.  324  the  wheel  consists  of  five  sections  bolted  at  the 
center  to  two  flanges  A  A,  called  spiders  or  centers.  The  sections 
are  held  together  at  the  rim  by  T-headed  links  B  which  are 
placed  in  cavities  in  the  rim.  These  links  and  those  of  the 
previous  figure  are  usually  placed  into  position  by  heating, 
the  length,  when  cooled,  being  enough  less  than  the  length 
between  ends  of  projection  that  the  extension  to  the  proper 
length  would  produce  the  force  desired  to  hold  these  together 
under  the  centrifugal  load. 

The  fly  wheel  is  used  at  times  as  a  driving  wheel  for  gears 
and  belts,  but  in  any  case  it  is  subject  to  stresses  of  a  complex 
nature.  These  will  be  examined  in  parts. 


394 


PUMPING 


Consider  the  rim  of  a  wheel  as  solid  and  free  from  spokes; 
on  each  square  inch  of  area  of  this  rim  there  is  a  centrifugal 
normal  force  of 


where  t=  thickness  in  inches, 
is)=  weight  of  one  cu.in., 
V  =  velocity  of  center  of  this  rim, 
r=  radius  of  center  of  rim  [%(r  ou 


in  ft. 


FIG.  324.—  Sectional  Fly  Wheel. 

Now  the  rim  undor  this  load  may  be,  considered  as  a  thin 
cylinder  subject  to  a  hydrostatic  pressure,  and  hence  the  formula; 
pd  =  2tS  holds. 

In  this  case 

p=F, 


5=  unit  stress. 


STEAM  END  DETAILS 


;95 


Hence 


twV* 
-- 


or  for  the  total  breadth 

—  V2i2b=tbS  =  T 
g 

When  T  is  the  total  pull  causing  rupture. 
The  unit  stress  5  becomes 

I2WV2 
o  =  . 

£ 

That  is,  the  unit  stress  in  a  rim  under  centrifugal  force  when 
rotated    at    the    velocity    V   is 
independent  of  the  radius  of  the 
wheel  and  size  of  the  rim,  and  is 
equal  to 

I2WV2 

- 

6 

For  cast  iron  w  =0.26  Ib.  per 
cu.in. 

S=-i\V2  approximately. 

This  stress  of  %wV2  produces 
an    elongation     in    the    radius 

wV2r 
of  !  —  —  ,and  if  spokes  are  used 


FIG.  325. — Rim  Shape. 


the  rim  will  take  the  exaggerated  form  shown  in  Fig.  325.  The 
spokes  will  be  stressed  by  the  rim  pull  and  in  turn  this  will  cause 
a  distortion,  as  shown,  so  that  at  points  A  A  there  will  be  no 
bending,  but  at  the  other  portions  of  the  rim  bending  forces 
will  result.  These  bending  forces  have  maximum  values  at  b 
and  c,  the  positive  maximum  being  at  c,  while  at  b  the  negative 
maximum  occurs. 

The  variation  of  these  stresses  has  been  investigated  theo- 


396  PUMPING  MACHINERY 

retically  by  Unwin  in  his  "Machine  Design,"  part  II,  pp.,  293 
301,  and  he  derives  the  following  formulae: 


FR    FR  cos   «- 

M  =  -  —  - 


2a       2          sin  OL 


_ 


2  sn  a 

(a-  <ft) 


2       sn  a 

Where        .F=pull  on  one  arm  in  pounds. 

M=  bending  moment  at  any  section  of  the  rim 

in  pound  feet. 

T  =  tension  at  any  section  of  the  rim  in  pounds. 
5=  shear  at  any  section  of  the  rim  in  Ibs. 
a=^  the  angle  between  the  arms  in  radians. 
G=  weight  of  metal  per  cu.ft.  in  Ibs. 
g  =32.16. 

A  =area  of  cross-section  of  rim  in  sq.ft. 
AI  =area  of  cross-section  of  spoke  in  sq.ft.  (mean). 
V=  velocity  of  center  of  rim  in  ft.  per  sec. 
R  =  radius  to  center  of  rim  in  ft. 
<£=  angle  in  radians  to  any  point  measured  from 
center  line  of  spoke. 

The  experimental  investigation  of  the  resistance  of  fly  wheels 
to  breaking  or  exploding,  as  it  has  been  called,  has  been  made 
by  Dean  Chas.  H.  Benjamin  of  Purdue.  In  these  investigations 
reported  to  the  American  Society  of  Mechanical  Engineers, 
(Vols.  20  and  22)  wheels  of  different  designs,  of  reduced  size, 

were  run  until  rupture  occurred.    At  this  point  the  speed  was 

72 
noted,    and   from   it   the   coefficient   —  ,   derived   above,   was 

10 

computed. 


STEAM   END  DETAILS  397 

In  the  case  of  a  1 5-inch  solid  wheel,  the  speed  at  rupture  was 

F2 
390  feet  per  second  (6000  R.P.M.)  and  the  value  of  —  =14,500. 

The  strength  of  the  cast  iron  from  which  this  wheel  was  made 
was  19,000  Ibs.  per  square  inch  in  direct  tension  and  39,000  was 
the  modulus  of  rupture.  With  a  flanged  jointed  wheel  of  24 
mches  diameter,  the  speed  at  rupture  was  190  feet  per  second, 

V2 
with  —  =3600  Ibs.  per  square  inch,  and  St  was  19,600.     With 

a  linked  joint  similar  to  Fig.  323  the  speed  at  rupture  was  305 

72 
feet  per  second  and  the  value  of  —  was  9300. 

From  these  and  other  tests  it  is  seen  that  the  solid  rim  wheel 
will  develop  slightly  less  than  its  tensile  strength,  while  the 
linked  wheel  will  have  about  two-thirds  the  strength  of  the 
solid  wheel,  and  the  flanged  jointed  fly  wheel  but  one-quarter 
the  strength. 

It  may  be  assumed  then  that  the  rim  is  designed  to  take 
the  centrifugal  tension. 

There  are  other  stresses  brought  on  the  rim  and  arms  when 
the  speed  changes.  Consider  a  wheel  in  which  the  arms  and 
rim  sections  are  not  united. 
When  the  engine  is  suddenly 
speeded  up  the  inertia  of  these 
sections  acts  as  a  load  on  the 
sections  and  bends  the  arms, 
as  shown  in  Fig.  326,  where  the 
rim  sections  are  joined.  This 
same  action  causes  the  rim  to 
take  the  form  shown.  This 
action  complicates  the  form- 
ula given  by  Unwin,  and  to 
cover  them  low  stress  should 

FIG.  326. — Effect  of  Bending. 

be  used,  and  then '  the  rim  is 

designed  for  the  centrifugal  load,  while  the  arms  are  designed 

to  take  the  bending  due  to  driving  or  to  inertia. 

The  twisting  moment  or  tangential  effort  diagram,  Fig.  283, 


398  PUMPING  MACHINERY 

gives  the  moment  for  which  the  arm  must  be  designed.  Let 
h  be  the  greatest  height  of  the  combined  tangential  effort  diagram 
which  shows  the  amount  of  effort  given  to  the  two  fly  wheels. 
The  moment  exerted  by  this  is 

JAX  scale  XL.P.  piston  area  X  crank  arm=M. 

This  is  the  maximum  moment  during  one  revolution  after 
the  engine  has  been  started,  but  if  steam  is  admitted  suddenly 
in  starting  the  pump,  the  entire  energy  from  the  piston  will 
be  transmitted  to  the  fly  wheels,  and  it  is  well  to  design  the  arms 
for  that  contingency.  This  represents  the  maximum  which 
could  ever  come  on  the  arms.  This  gives 

nd2 

Mmax=i—  4.P.  />boiierX  crank  arm. 
4 

This  moment  is  divided  between  m  arms,  hence 


m  32  ' 

for  the  elliptical  arms  of  diameters  b  and  d. 

This  is  really  the  moment  and  size  at  the  center  of  the  shaft 
and  the  arms  are  tapered  on  each  diameter  to  the  rim  to  two- 
thirds  these  dimensions. 

The  rim  of  a  solid  wheel  is  sufficiently  strong,  provided  V 
be  determined  by  the  formula 

V2 

—  <  safe  stress. 

10 

There  is  no  way  of  making  this  stronger  if  the  material 
is  fixed.  For  high  speeds  steel  may  be  used  to  increase  the 
right  hand  side  and  so  permit  a  higher  velocity. 

The  links  used  to  join  the  sections  are  subject  to  a  load 

fze'F2Xarea  of  section. 

Hence,  if  there  are  n  sections  of  link,  each  of  area  a  at  the 
section, 

=naSt. 


STEAM  END  DETAILS  399 

From  this  a  can  be  found,  and  from  it  the  cross-section 
dimensions  of  the  link. 

The  bolts  at  the  center  of  the  wheel,  as  in  Fig.  323  or  324, 
should  be  large  enough,  if  possible,  to  carry  the  section  if  the 
links  should  break.  To  find  the  radial  force  produced  b}'  a 
section  of  the  wheel,  it  is  necessary  to  find  the  center  of  gravity 
of  the  section  since 

CF-^F2 

C  -F'-g    R> 

where  W  is  the  weight  of  the  rotating  body,  V  the  velocity  of 
the  center  of  gravity  and  R  the  radius  to  that  point. 

The  center  of  gravity  of  a  section  of  a  ring  of  angular  extent 
a  and  of  thickness  t  and  mean  diameter  D,  is 


This  then  gives  the  centrifugal  force  from  the  rim  of  a  section, 
to  which  may  be  added  that  of  the  arm  by  considering  the 
center  of  gravity  to  be  at  the  center  of  the  length  of  the  arm. 
Calling  these  forces  JFc.f.,  the  following  gives  the  size  of  the  n' 
bolts  in  tension  or  the  n"  bolts  in  double  shear  to  resist  this, 


c,. 


The  center  discs  for  the  built  up  wheel  and  the  hub  of  the 
split  wheel  are  usually  of  empirical  design,  and  in  the  figure  the 
unit  has  been  taken  equal  to  x/rim  area. 

The  fly  wheels  should  be  keyed  to  the  shaft.   The  force  coming 

M- 
max 

on  the  key  is,  -  —  ,    f,    ,.  -  . 
'  £  shaft  diam. 

M 


f 
shaft  diam. 

/=  length  of  key  in  inches, 
b  =  breadth  of  key  in  inches, 
7  --  depth  of  key  in  inches. 


400 


PUMPING  MACHINERY 


The  balancing  of  fly  wheels  of  pumps  is  a  simple  matter,  for 
the  speeds  are  low,  and  all  that  is  necessary  is  to  have  the  wheel 
in  static  balance.  The  balancing  of  the  reciprocating  parts 
is  not  important  in  the  slow-running  pumps.  For  a  detailed 
discussion  of  balancing  of  the  reciprocating  and  rotating  parts 
the  reader  is  referred  to  "The  Balancing  of  Engines,"  by  W. 
E.  Dalby. 

FRAMES 

The  frames  of  pumps  are  of  different  forms.  Fig.  327  shows 
a  type  in  which  the  A  frames  supporting  the  cylinders  are 


FIG.  32% — Frame  for  Vertical  Engine: 

carried  on  bed  plates.  The  bed  plate  is  carried  on  the  air  chanv 
bers  on  onp  side  and  on  masonry  on  the  other.  The  design  ol 
such  structures  is  from  experience.  The  thickness  of  the  metal 
usually  being  about  one  inch  and  the  sizes  of  the  columns  are 


STEAM  END  DETAILS  401 

fixed  by  what  has  been  done.  The  load  coming  on  the  frame, 
if  inertia  and  friction  are  neglected,  is  equal  to  the  net  steam 
pressure  multiplied  by  the  area  of  the  piston.  This  force  has  to 
be  transmitted  down  in  case  of  the  upstroke,  and  on  the  down- 
stroke  the  pressure  is  upward  and  is  resisted  by  the  weight  of 
this  section  of  the  engine  and  the  weight  of  the  masonry  when 
the  pump  is  carried  below  on  the  foundation.  In  the  case  of  a 
single-acting  pump  the  major  part  of  the  force  on  the  upstroke 
is  transmitted  to  the  main  bearing  and  this  causes  a  bend  in 
the  frame  of  the  overhung  crank  engine.  This  bending  moment 
is  equal  to 

PX  (distance  a,  Fig.  316). 

The  moment  is  resisted  by  the  frame.  On  the  downstroke 
the  pressure  from  the  steam  is  balanced  by  the  water  pressure, 
which  is  central,  and  so  there  is  no  tendency  to  bend.  To  cut 
down  vibrations  in  the  frame  cross  pieces  or  knees  are  put  in 
at  frequent  intervals. 

Fig.  346  shows  a  type  of  frame  in  which  cast  iron  columns 
are  used  to  carry  the  weight  to  the  lower  foundations,  while  in 
Fig.  349  the  weight  is  carried  by  the  air  chambers  entirely, 
a  method  which  gives  good  results,  as  it  clears  space  around  the 
base,  facilitating  repair,  examination  or  operation.  By  examin- 
ing the  various  figures  of  the  horizontal  and  vertical  pumps,  the 
methods  of  .general  practice  may  be  seen. 

The  operating  floors  are  carried  at  different  levels  from 
beams  bolted  to  the  frames.  These  should  be  designed  to  carry 
about  two  hundred  pounds  per  square  foot. 

The  guide  surfaces  are  carried  in  the  frame,  and  if  not  a 
portion  of  the  main  casting  the  connections  should  be  made 
to  carry  the  load 

P' 
n 

The  foundations  are  designed  from  experience.  There  are 
two  principles  to  consider:  First,  the  foundation  should  be  large 
enough  to  properly  distribute  and  carry  the  load,  and  second, 


402  PUMPING  MACHINERY 

there  should  be  sufficient  mass  to  reduce  vibrations  of  the 
machine. 

The  area  of  the  base  of  the  foundation  should  be  such  that 
the  weight  carried  from  the  engine  and  foundation  is  3000  to 
6000  pounds  per  square  foot.  The  first  value  is  for  good  clay 
soil,  the  latter  for  a  compact  gravel.  To  get  this  it  may  be 
necessary  to  spread  the  foundation,  and  to  do  this  a  batter  of 
about  one  foot  in  two  or  three  feet  of  height  will  be  found  proper 
to  develop  the  strength  of  the  projecting  part.  The  foundations 
may  be  made  of  concrete  of  proportions:  one  part  cement, 
two  parts  sand  and  five  parts  broken  stone;  or  hard  burned 
bricks,  set  in  cement  mortar,  may  be  used.  These  brick  should 
be  thoroughly  wet  when  laid  and  the  joints  should  be  grouted 
with  a  watery  mixture  of  cement  mortar. 

The  amount  of  material  to  be  placed  in  the  foundation  to 
reduce  vibrations  is  a  matter  of  experience.  In  a  slow  running 
machine  there  is  practically  no  need  of  heavy  foundations, 
and  when  the  engine  is  even  of  high  speed  and  properly  balanced, 
there  is  no  need  of  making  the  foundation  very  thick.  A  thick- 
ness of  three  or  four  feet  is  sufficient  for  horizontal  machines, 
and  for  vertical  machines  the  thickness  will  be  determined  by 
the  amount  of  height  necessary  to  give  the  proper  spread  to 
the  base. 

When  the  engine  is  to  be  placed  over  a  very  soft  soil,  it  is 
necessary  to  drive  piles  for  the  support  of  the  masonry.  These 
are  laid  out  so  as  to  support  a  given  load,  say  forty  tons  apiece, 
and  when  driving,  the  pile  is  driven  until  the  amount  of  pene- 
tration under  a  given  blow  of  the  hammer  shows  that  it  will 
carry  the  desired  load.  If  this  load  can  not  be  obtained  the 
piles  must  be  placed  close  together.  At  times  pier  holes  may 
be  sunk  over  the  engine  base  area  to  good  soil  or  rock  and 
the  main  foundation  can  be  carried  on  the  piers  built  in 
them. 

A  formula  recommended  by  Baker  for  bearing  power  of  a 
pile  is 


P  =  100  (VWh  +  (5<><*)2~5od), 


STEAM  END  DETAILS  403 

where     P  =  supporting  pressure  in  tons, 
W  =  weight  of  ram  in  tons, 
A=height  of  fall  in  feet, 
d= penetration  of  pile  in  feet  at  last  blow. 

The  piles  should  be  of  good  quality,  of  sound  white  oak,  not 
less  than  10  inches  in  diameter  at  the  smaller  end  and  14  inches 
at  the  larger.  They  should  be  straight  grained  and  have  all 
bark  removed.  After  driving  they  are  cut  off  below  the  per- 
manent water  line,  and  then  concrete  is  put  around  the  end, 
making  a  solid  bed. 

SPECIAL  STEAM  PIPING  AND  VALVES. 

The  steam  piping  used  on  the  engine  should  be  designed  so 
that  the  velocity  of  the  steam  is  6000  feet  per  minute.  This  is  a 
simple  method  and  gives  good  results.  There  is  a  method  in 
which  the  pipe  is  designed  to  give  a  certain  discharge  when  the 
drop  in  pressure  is  assumed.  The  formula  developed  by  Prof. 
Carpenter  for  this  (A.  S.  M.  E.,  vol.  xx,  p.  342)  is 


/       3 
V    f  d' 


2o.663     V     '   r/Af*1 

where    p  =loss  of  pressure  in  Ibs.  per  sq.in., 
k  =  a  constant  =0.0027, 
d'  =  diameter  of  pipe  in  inches, 
ze>'=flow  of  steam  per  minute  in  pounds, 
L  =  length  of  pipe  in  feet, 
D=  weight  of  one  cu.ft.  of  steam. 

This  may  be  used  when  long  pipes  are  employed,  but  the 
rule  above  for  the  area  will  give  satisfactory  results  in  pumping 
stations,  and  even  8000  feet  per  minute  may  be  used  when  neces- 
sary for  large  pipes. 

The  steam  lines  should  be  made  of  heavy  full  weight  pipe 
and  tested  to  stand  250  pounds  per  square  inch.  In  this  work 
most  of  the  pipes  are  large  and  the  joints  are  made  by  flanges. 


404 


PUMPING  MACHINERY 


Fig.  328  shows  the  proportion  of  the  standard  screwed  flange, 
while  Fig.  329  shows  a  similar  flange  with  a  caulking  recess  on 


FIG.  328.— Flange. 


FIG.  329. — Screwed  Flange 
with  Caulking  Recess. 


FIG.  330. — Shrunk  Flange. 


FIG.  331, — Rolled  Joint. 


the  back  which  may  be  filled  with  a  metal  for  caulking.  This 
is  not  always  advisable.  These  two  figures  show  the  pipe  screwed 
on  the  flange.  This  is  done  so  that 
the  pipe  projects  a  slight  distance 
beyond  the  flange  when  it  is  tight- 
ened to  its  full  extent,  and  then  the 
projection  is  turned  off  flush  with  the 
flange.  The  thread  of  the  pipe  is 
cut  deep  to  accomplish  this  result. 


FIG.  332.— Welded  Flange.  FIG.  333.-— Globe  Valve. 

Flanges  are  now  shrunk  on  the  pipe,  Fig.  330,  after  which 
the  end  of  the  pipe  is  peened  into  a  cavity  left  in  the  flange,  and 


STEAM  END  DETAILS 


405 


then  the  end  is  turned  flush  with  the  flange.  The  Crane  Company 
sometimes  roll  the  pipe  into  grooves  turned  in  the  flange,  as 
shown  in  Fig.  331.  This  operation  is  similar  to  the  expanding 
of  boiler  tubes  into  the  heads  of  boilers. 

The  welded  flange,  Fig.  332,  is  a  new  form  of  flange  conneo 


M-. 


FIG.  334. — Ludlow  Gate  Valve. 

tion  for  joining  pipes,  and  according  to  the  manufacturer  it  is 
meeting  with  favor  for  high  pressure  work. 

The  dimension  of  these  flanges,  the  number  and  sizes  of  the 
bolts,  the  size  of  various  fittings,  bends  and  other  specials  for 
pipe  work,  are  to  be  found  in  the  catalogues  and  on  the  dimension 
sheets  of  Crane,  Walworth,  or  the  other  manufacturers  of  pipe 
fittings. 


406 


PUMPING   MACHINERY 


The  valves  used  are  of  the  globe  or  gate  types.  The  globe 
valve  shown  in  Fig.  333  is  of  the  general  form  with  an  outside 
yoke,  and  the  type  of  gate  valve  shown  in  Fig.  334  is 
often  employed.  The  dimensions  of  these  valves  are  found  in 
tables  furnished  by  the  manufacturers,  so  that  in  laying  out 
work  the  engineer  may  know  how  much  to  allow  for  these  fittings. 


FIG.  335. — Crane  Gate  Valve. 

The  valves  may  have  screw  ends,  Fig.  333,  or  flange  ends,  Fig. 
334,  while  for  water  pipe  bell  ends  are  used.  The  large  gate 
valves,  Fig.  335,  used  on  water  lines  are  built  with  gearing 
to  turn  the  spindle. 

CONDENSERS. 

The  condensers  used  are  either  jet  or  surface  condensers. 
Wht,n  the  jet  condensers  are  used,  the  volume  of  the  con- 
denser head  should  be  one  third  of  the  volume  of  the  low 


STEAM   END  DETAILS  407 

pressure  cylinder.  The  pipe  through  which  the  water  enters  the 
head  should  be  such  that  the  velocity  is  2  \  feet  per  second.  That 
is, 


_ 
62.5  X  2  £' 

^4  =net  area  pipe  in  sq.in. 
Wcw.  =  water  required  per  sec.  in  Ibs. 

The  surface  of  the  surface  condenser  should  be  determined 
so  that  the  condensation  may  be  accomplished  by  the  trans- 
mission of  500  B.T.U.  per  hour  per  sq.  ft.  of  surface  per 
degree  difference  in  temperature. 

This  gives 

W(H-qQ') 


5ooU- 


where 

5=  surf  ace  in  square  feet, 
W  =  weight  of  steam  condensed  per  hour, 
H  =  heat  content  of  exhaust  steam 

=q  +  xr  or  approximately  q+r, 
ts  =  temperature  of  steam  in  condenser, 
tQ=  temperature  of  condensing  water  leaving  condenser, 
ti  =  temperature  of  condensing  water  entering  condenser. 

In  the  Journal  of  the  American  Society  of  Mechanical  Engi- 
neers for  November,  1910,  Mr.  Geo.  A.  Orrok  gives  a  resume 
of  a  greater  number  of  experiments  and  theoretical  papers  on 
the  transmission  of  heat  by  condenser  tubes  as  well  as  the  results 
of  a  great  number  of  experiments  of  his  own  and  the  deductions 
drawn  from  all  of  these.  In  this  paper  Orrok  derives  the 
equation, 


S=  square  feet  of  condenser  surface. 
W  =  steam  condensed  per  hour  in  pounds. 


408  PUMPING  MACHINERY 

G  =  condensing  water  per  pound  of  steam  in  pounds. 

ts  =  temperature  of  steam. 

ti  =  temperature  of  condensing  water  at  inlet. 

to  =  temperature  of  condensing  water  at  outlet. 

c=  cleanness  factor,  varying  from  i.oo  with  clean  tubes 

to  0.5  with  dirty  tubes. 
fi= material  coefficient. 

=  r.oo  for  copper. 

=0.98  for  Admiralty  tubes. 

=-0.97  for  Admiralty  aluminium  lined. 

=  0.92  for  Admiralty  oxidized  (black). 

=  0.87  for  aluminium  bronze. 

=  0.80  for  cupro-nickel. 

=0.79  for  tin. 

=0.75  for  zinc. 

=  0.74  for  Monel  metal. 

=0.63  for  Shelby  steel. 

=0.55  for  Admiralty  badly  corroded. 

=0.47  for  Admiralty  vulcanized  inside. 

=0.25  for  glass. 

^0.17  for  Admiralty  vulcanized  both  sides. 

p-~-  ratio  of  the  steam  pressure  corresponding  to  the  tem- 

p 
perature  to  the  total  vacuum  pressure  =— *. 

-*  c 

Vw=  velocity  of  water  in  tubes  in  feet  per  second. 
Now      0=mean  temperature  difference  in  condenser 


r 


In  practice,  when  ^  =  70°  F.,  ^=90°  F.,  t8=--g8°  F.,  0  =  16.5. 
If  there  is  a  28-inch  vacuum,  and  the  tubes  are  medium  clean 
Admiralty  tubes  in  which  the  water  is  flowing  at  4  feet  per 
second,  the  following  value  is  found  for  S  for  each  1000  pounds 
of  cooling  water  per  hour: 


•9437 


al  Pressure 

q.  in. 

Sq.  ft. 
per  i  H.P. 
.  I.7T 

I    ^7 

*  O/ 
.  I    ^O 

.1  .43 

I    37 

.1.30 

STEAM  END    DETAILS  .      409 

Sometimes  the  surface  is  determined  by  the  I.H.P.  of  the 
engine.  The  following  table  is  taken  from  Peabody's  "  Ther- 
modynamics of  the  Steam  Engine": 


20 
15 
12* 

10 

8 
6 

The  air  pump  of  the  jet  condenser  is  made  sufficiently  large 
to  care  for  the  condensed  steam,  the  condensing  water  and  the 
air.  For  this  purpose  it  is  made  so  that  its  displacement  is 
about  forty  times  the  volume  of  the  condensed  steam.  For  wet 
air  pumps  used  with  surface  condensers,  the  pump  handles  only 
the  condensed  steam  and  air,  and  therefore  the  volume  displaced 
is  reduced  to  one-half  of  the  former  amount  or  twenty  times 
the  volume  of  the  condensed  steam.  A  better  method  of  deter- 
mining the  volume  will  be  given  in  Chapter  XI. 

The  quantity  of  cooling  water  to  be  supplied  these  condensers 
and  to  be  cared  for  by  the  pumps  is  given  by  the  equation: 

W(H-qo') 

*>-?<      ' 
Wc.w.  =  amount  of  cooling  water  to  be  used  with  W  Ibs. 

of  steam; 
H  =  heat  content  of  exhaust  steam  =q+xr  assumed 

to  be  q  +  r; 

<?o/i=heat  of  liquid  of  cond3nsed  steam; 
<7o=heat  of  liquid  of  condensing  water  leaving; 
<?j=heat  of  liquid  of  condensing  water  entering. 


CHAPTER  IX 
TEST  OF  PUMPING  ENGINES 

DUTY  TRIAL 

ENGINES  are  usually  tested  for  steam  consumption;  that 
is,  the  number  of  pounds  of  steam  per  horse-power-hour,  but 
with  pumps,  their  duty  is  usually  determined.  Duty  is  the 
number  of  foot  pounds  of  useful  work  done  by  one  thousand 
pounds  of  dry  steam  or  one  million  British  thermal  units  used 
by  the  pump.  The  problem,  then,  is  to  find  the  number  of  foot 
pounds  of  work  done  by  the  pump,  and  the  number  of  heat  units 
used  by  the  pump.  Duty  meant  originally  the  amount  of  work 
per  hundred  pounds  of  coal,  so  that  different  pumps  could  be 
compared.  As  coals  varied  this  did  not  have  a  definite  meaning, 
and  so  a  more  definite  unit  was  desired.  The  value  1,000,060 
was  taken  because  on  an  average  one  pound  of  coal  gives  10,000 
B.T.U.  to  the  water  in  the  boiler,  and  hence  1,000,000  B.T.U, 
is  equivalent  to  an  average  of  TOO  pounds  of  coal,  while  1,000 
pounds  of  steam  can  be  assumed  to  be  evaporated  by  this  coal,  as 
each  pound  of  water  requires  about  1,000  B.T.U.  to  evaporate  it. 

In  considering  the  useful  work  done,  the  work  necessary 
to  bring  the  water  to  the  pump  and  that  necessary  to  send  it 
from  the  pump  to  the  reservoir  must  be  included.  The  work 
necessary  to  overcome  the  friction  in  the  pump  itself,  i.e., 
through  valves  and  losses  of  sudden  contraction,  etc.,  is  not 
included,  for  these  are  due  to  the  pump.  With  this  in  mind  it 
is  seen  that  the  useful  work  is 


W  2g 

where  W  is  total  useful  work;  w  the  weight  of  one  cubic  foot  of 
water;  Q  the  total  number  of  cubic  feet  pumped;  pi  the  pres- 


410 


TEST  OF  PUMPING   ENGINES  411 

sure  head  in  the  force  main  in  pounds  per  square  inch;  p2  the 
vacuum  in  the  suction  pipe  in  pounds  per  square  inch,  and  h 
the  distance  in  feet  between  the  force  gauge  and  the  suction 
gauge;  or 

W  =  Q  (  i44p  i  +  wh  +  144^2  ) 


where  the  symbols  mean  the  same  as  before. 

The  value  of  w  can  be  found  from  the  temperature  of  water 
pumped,  and  in  important  cases  a  known  volume  of  water 
should  be  weighed  to  find  the  effect  of  suspended  matter. 

The  quantity  of  water  is  usually  determined  in  one  of  five 
ways:  first,  by  displacement;  second,  by  weirs;  third,  by 
meters;  fourth,  by  weighing;  fifth,  by  measuring. 

First,  the  diameter  of  the  plunger  is  measured  and  also  the 
length  of  the  stroke.  These,  with  the  number  of  the  strokes, 
give  the  displacement,  or  the  quantity  of  water  pumped  if  there 
is  no  leakage,  or  slip,  as  it  is  sometimes  called.  This  slip  occurs 
around  the  piston  or  plunger  in  an  inside  packed  piston;  but  with 
the  method  of  outside  packing  used  to-day  this  leakage  is 
nothing  and  the  only  other  place  for  leakage  to  occur  is  past  the 
valves.  In  order  to  determine  the  leakage  around  the  piston 
when  inside  packed,  and  the  valves  when  seated,  standard 
methods  are  used.  (See  Carpenter,  "  Experimental  Engineering," 
page  560,  and  Kent,  "Pocket  Book,"  page  611.)  The  leakage 
past  the  valves  in  seating  cannot  be  ascertained,  except  by 
determining  Q  by  one  of  the  other  methods. 

In  some  pumps  the  stroke  of  the  plunger  is  not  definite, 
and  hence  the  stroke  should  be  measured  continuously.  If 
conditions  are  nearly  constant  the  stroke  will  not  vary  much,  and 
hence  readings  of  length  taken  at  regular  intervals  are  made. 
This  is  not  absolutely  correct,  and  hence  some  form  of  registering 
apparatus  should  be  employed.  (See  A.S.M.E.,  vol.  12,  page 
981.) 

Having  now  the  areas,  stroke  and  leakage,  the  amount 
pumped  may  be  computed 


100 


412  PUMPING 

Second,  for  this  method  the  constants  for  the  weir  and 
also  the  hook  gauge  must  be  known.  Attempt  should  be  made 
to  reduce  the  fluctuation  of  the  water  surface,  and  the  gauge 
should  be  read  in  a  box  on  one  side  of  the  flume. 

Third,  since  the  introduction  of  the  Venturi  meter  this 
instrument  can  be  used  to  advantage  in  pump  tests,  as  its  error 
is  small.  The  Pi  tot  tube  also  serves  as  a  most  accurate  method 
of  determining  the  quantity  if  a  traverse  is  made  by  a  tube  of 
small  size  pointing  in  direction  of  the  current  when  the  static 
pressure  is  taken  from  holes  in  the  pipe  walls. 

Fourth,  for  small  pumps  the  water  pumped  can  be  actually 
weighed,  thus  giving  the  exact  amount  pumped. 

Fifth,  cisterns  and  reservoirs  may  be  used  when  their 
contents  are  known,  but  here  allowance  must  be  made  for  the 
evaporation  from  the  surface  and  the  effect  of  wind. 

The  quantity  p\  is  determined  by  means  of  a  pressure  gauge 
which  must  be  tested.  (Be  sure  to  take  maker's  number  of  all 
instruments  used  and  position  where  used.)  Since  the  pressure 
in  the  force  main  varies  considerably,  the  gauge  cock  is  throttled 
off,  or  an  air  chamber  is  introduced  between  the  gauge  and  the 
force  main  which  will  reduce  *the  fluctuations.  p2  is  usually 
determined  by  a  mercury  tube,  and  hence  must  be  reduced 
from  inches  of  mercury  to  pounds  per  square  inch.  This  pressure 
is  greater  than  the  head  between  the  points  where  measured  and 
the  water  in  the  forebay,  as  it  is  equal  to  the  velocity  head  and 
the  head  of  friction  loss,  in  addition  to  the  lift.  This  should  be 
included,  as  it  is  not  on  account  of  the  pump  that  such  losses 
occur,  but  on  account  of  the  location  of  the  pump  demanding 
a  certain  length  of  pipe. 

The  difference  of  level  between  the  suction  gauge  and  force 
gauge  must  be  taken,  because  the  water  is  raised  through  this 
distance,  but  the  distance  is  not  given  by  either  of  the  pressure 
readings. 

Having  these,  the  useful  work  can  be  found.  If  indicator 
cards  are  taken  from  the  cylinder,  they  give  the  work  done  on 
the  water,  and  the  difference  between  this  and  the  useful  work 


TEST  OF  PUMPING  ENGINES  413 

just  determined,  would  give  the  losses  due  to  leakage  and  the 
friction  of  the  water  in  the  pump. 

The  readings  of  pressure  and  the  revolution  counter  are 
taken  at  regular  intervals,  and  the  average  result  used. 

To  determine  the  quantity  of  heat  used  by  the  engine,  the 
quantity  of  heat  in  each  pound  of  steam  used  in  the  engine 
must  be  found  as  well  as  the  total  number  of  pounds  of  steam 
used.  The  best  way  to  determine  the  quantity  of  steam  used  is 
to  condense  the  steam  in  a  surface  condenser  and  weigh  it.  To 
this  must  be  added  the  amount  consumed  in  the  steam  jackets 
and  the  superheaters. 

The  water  coming  from  the  jackets  and  superheaters  is  at 
such  a  temperature  that  it  would  partially  evaporate  if  allowed 
to  discharge  into  the  air  before  being  cooled.  There  are  two 
methods  of  weighing  this:  one  consists  of  allowing  the  hot  water 
to  discharge  into  a  tank  partially  rilled  with  cold  water;  the 
other  is  to  conduct  it  through  a  coil  surrounded  by  cold  water 
and  discharge  into  tanks. 

As  condensers  attached  to  pumping  engines  are  often  jet 
condensers  the  above  method  is  impossible,  and  so  the  following 
method  must  be  used:  A  number  of  boilers  sufficient  to  run 
the  pump  to  be  tested  are  separated  from  the  other  boilers  and 
all  cross  connections  of  the  water  or  steam  pipes  are  broken 
and  blanked  off.  Then  no  steam  may  be  used  for  any  other 
purpose  and  no  water  can  be  brought  into  the  boilers,  except 
by  the  one  feed  pump,  which  draws  its  supply  from  a  tank,  the 
water  in  which  has  been  weighed. 

Now  if  there  is  any  water  collected  in  a  separator  between 
the  boilers  and  the  engine,  or  used  for  any  purpose  whatsoever, 
other  than  for  the  engine,  it  must  be  carefully  determined. 

Suppose  W  pounds  of  water  are  pumped  into  the  boiler, 
w  pounds  used  in  a  calorimeter  to  determine  the  quality  of  the 
steam,  w'  pounds  taken  from  the  separator,  and  w"  pounds  used 
in  the  jackets  and  reheaters.  The  amount  of  water  used  then  by 
the  main  pumping  engine  is  W  —  w-w'  pounds.  Care  must 
be  exercised  to  investigate  the  piping  to  see  that  there  is  no 


414  PUMPING  MACHINERY 

chance  of  leakage  of  the  water  or  steam  in  any  way,  and  all 
joints  into  other  lines  should  be  broken  and  blanked  off.  A 
leakage  test  of  several  hours'  duration  should  be  made  at  the 
end  of  the  test,  with  the  main  engine  shut  off  to  determine  the 
loss  from  the  boilers  and  piping.  The  quantity  so  found  is 
subtracted  from  W. 

By  taking  readings  of  the  steam  gauge,  barometer,  and  on 
the  calorimeter,  the  absolute  pressure  and  the  quality  of  the 
steam  can  be  determined,  and  hence  the  quantity  of  heat  above 
32°  F.  supplied  per  pound.  The  total  heat  supplied  above 
32°  F.  is  then 

H  =  (W-w-w')(q+xr) 
where 

<7=heat  of  the  liquid; 

x  =  quality  of  steam; 

r=heat  of  vaporization. 

For  superheated  steam  q+xr  becomes  q  +  r  +  I  cpdt. 

In  general,  the  steam  jackets  and  reheaters  drain  through  a 
trap  into  the  boiler  feed,  and  hence  return  some  heat  to  the 
boiler.  To  determine  this  the  weight  of  the  water  from  them 
is  found  as  before  described  and  the  temperature  of  that  water 
before  it  is  cooled.  The  temperature  is  found  by  placing  a 
thermometer  cup  in  the  pipe  and  taking  temperature  readings 
on  a  thermometer  placed  therein.  The  heat  then  returned  is 

»'V 

The  feed  pump,  under  actual  running  conditions,  takes 
water  from  the  hot  well,  hence  the  temperature  of  this  must 
be  observed  and  the  heat  supplied  by  this  source  determined. 
If  the  temperature  be  £3,  the  quantity  returned  is 

(W-w-w'-w")q3 
The  total  heat  chargeable  to  the  engine  is 


then 

Work 

—  —  i,  000,000=  Duty 


TEST  OF  PUMPING   ENGINES  415 

If  the  duty  to  be  found  is  on  the  basis  of  1,000  pounds  of  dry 
steam,  the  steam  used  must  be  expressed  in  terms  of  equivalent 
dry  steam.  The  equivalent  dry  steam  is  given  by 

(W-w-w')(q+xr-q3)    ., 

W£=- : TT-: ,  if  saturated, 

(q+r-qs) 

or 

/  /-/sup.  \ 

(W -w-w'}\Q+r+  I         cdt-qzj 

We  --  _^_       _L   if  superheated. 

(q+r-q3) 

Work 
Then  duty  =    „,      1,000. 

There  are  other  quantities  which  are  reported.  The  mechan- 
ical efficiency  of  the  pump  is  determined  by  taking  indicator 
cards  from  the  water  and  steam  cylinders  and  determining  the 
horse-power  of  each  end,  and  the  ratio  of  these  gives  the  mechan- 
ical efficiency.  The  first  subtracted  from  the  second  gives  the 
frictional  loss. 

Call  the  horse-power  of  the  steam  end  I.H.P.  and  of  the 

'  D  H  P 

water  end  D.H.P.     Then  mechanical  efficiency  ~  T  TT  W"- 

l.rl.r . 

I.H.P.  -D.H.P.  =H.P.  of  engine  friction. 

The  quantity  of  steam  used  per  hour,  divided  by  the  I.H.P., 
is  the  steam  per  I.H.P.  hour,  a  result  useful  for  comparison  with 
other  kinds  of  engines.  This  is 

W  —  w  —  w' 

=  Steam  per  I.H.P.  hour. 


I.H.P. 

or  better 

We 

prrp-  =Dry  steam  per  I.H.P.  hour.  . 

The  quality  of  the  steam,  or  in  other  words  the  amount 
of  condensation,  in  each  cylinder  at  any  time  after  cut-off  in 
that  cylinder  can  be  determined,  first,  by  finding  the  volume 
(including  clearance  volume)  and  pressure  at  that  point  from 


416  PUMPING  MACHINERY 

the  indicator  cards,  and  second,  from  the  table  of  the  properties 
of  steam,  the  volume  occupied  by  one  pound  of  steam  at  this 
pressure.  From  the  quantity  of  steam  used  in  the  cylinder,  the 
number  of  pounds  per  stroke  is  found;  and  on  adding  to  this 
the  steam  in  the  clearance  space  the  total  steam  in  the  cylinder 
may  be  ascertained.  The  clearance  weight  is  approximated  by 
considering  the  steam  dry  at  compression  and  then  finding 
weight  by  the  formula: 

Volume 


Wt 


Specific  Volume' 


The  product  of  the  specific  volume  and  the  total  weight  per 
stroke  gives  the  volume  which  should  be  occupied.  The  ratio 
of  the  actual  volume  at  the  point  to  this  amount  represents  the 
percentage  steam  present  at  that  point,  the  remainder  being 
water,  the  volume  of  which  has  been  neglected.  In  this  way 
the  initial  condensation  may  be  determined. 

The  amount  of  heat  removed  by  the  surface  condenser  is 
W'(q'-q"}+W"q'",  where  W  is  the  weight  of  the  condensing 
water,  and  q'  is  the  heat  of  the  liquid  leaving  the  condenser  and 
q"  the  heat  of  the  liquid  in  the  entering  condensing  water.  W" 
is  the  weight  of  the  condensate  and  q'"  its  heat  of  the  liquid. 

This  quantity  of  heat  removed  by  the  condenser  may  be 
called  #';  then 

H-H'-  Work  =  Radiation, 

Radiation     ^  ,       ,          ,.  .. 

or  100 — =  %  loss  by  radiation. 

H 

The  following  precautions  are  to  be  noted  in  making  obser- 
vations : 

First,  carefully  arrange  observation  sheets  and  become 
familiar  with  readings  before  beginning  test.  . 

Second,  after  noting  observations  make  a  check  observation 
when  possible. 

Third,  keep  a  vigilant  watch  for  leaks  of  any  kind. 

Fourth,  note  everything  in  book  that  happens  out  of  the 
ordinary. 


TEST  OF  PUMPING    ENGINES  417 

Fifth,  make  note  of  number  and  position  of  all  instruments 
used. 

Sixth,  calibrate  and  check  all  instruments  and  measure- 
ments. 

Seventh,  in  weighing  coal  or  water,  take  the  weight  of  vessel 
empty  and  full,  together  with  the  time,  in  each  case. 

Eighth,  note  the  time  of  all  observations,  except  when 
several  different  ones  are  made  in  quick  succession,  then  note 
the  time  of  the  first  only,  but  take  them  always  in  the  same 
order. 

Ninth,  always  have  instruments  ready  for  next  operation 
before  leaving  them. 

The  data  and  results  called  for  in  duty  trials  are  given  in 
Kent's  Pocket-book,   and  the  following  additional  results  and 
data  are  also  advisable  for  investigation  of  the  pump  action: 
Plunger  displacement  per  stroke. 
Amount  pumped,  cubic  feet. 
Quantity  pumped,  computed  from  displacement. 
Quantity  of  water  used  in  condenser. 
Quantity  of  heat  absorbed  by  condenser. 
Quantity  of  heat  absorbed  by  condenser,  per  square  foot 

of  surface. 

Percentage  loss  by  radiation. 

These  tests  are  usually  made  to  find  whether  or  not  the 
guaranteed  duty  is  met,  but  an  investigation  in  the  same  manner 
may  be  made  to  find  whether  or  not  a  given  device  or  arrange- 
ment is  of  value.  For  instance,  in  many  cases  it  is  found  that 
the  use  of  the  reheating  coils  in  the  receivers  are  not  a  source 
of  gain  and  at  times  a  test  without  certain  jackets  will  give 
better  results  than  a  test  made  with  the  jackets  in  service. 

Superheated  steam  can  be  used  and  its  effect  determined 
in  this  manner.  The  use  of  sufficient  superheat  to  cut  down 
the  initial  condensation  and  give  dry  steam  at  cut-off  will  pay, 
but  beyond  this  the  expense  in  making  the  superheated  steam 
will  be  more  than  the  gain  in  the  engine.  The  only  reason  why 
it  should  be  of  value  when  the  initial  condensation  is  eliminated, 
is  the  fact  that  when  the  steam  is  dry  or  superheated  the  walls 


418 


PUMPING  MACHINERY 


do  not  take  up  as  much  heat  as  when  covered  with  a  film  of 

moisture. 

One  reason  why  superheated  steam  should  not  add  to  the 

efficiency,  the  amount  apparently  shown  by  the  increase  of  the 
upper  temperature  T\  of  the  expression 
for  the  Carnot  efficiency, 


-"  —  T  y 

1  1 

is  the  fact  that  this  heat  is  not  supplied 
at  a  fixed  temperature  7^,  but  at  a  grad- 
ually increasing  temperature,  and  hence 
although  the  efficiency  of  the  cycle  is 
increased  a  slight  amount  the  increase 
is  very  small.  This  may  be  shown  by 
Fig.  336,  which  is  a  T<j>  diagram.  In 

the  figure,  i,  2, 3, 4  represents  the  cycle  of  an  engine  using  wet 
steam,  while  i,  2,  3^,4'  is  one  for  an  engine  supplied  with  super- 
heated steam.  The  efficiencies  are 


5  6        G 

FIG.  336. — Entropy 
Diagram. 


and 


~ 


5,1,2,3,6 


i,  2.  3'.  4' 


1,2,7,4' 


1274 
51276 


is  slightly  greater  than 


T    ^    Q     A 

'    ' 


so  that  £sa   and  E 


sat 


su 


are  almost  the  same. 

The  use  of  jackets  on  the  cylinder  heads  and  barrels  is 
recommended  although  tests  do  not  always  show  an  increase 
of  economy. 

The  overall  efficiency  of  the  engine  is  given  by  dividing 
the  duty  per  1,000,000,  B.T.U.  by  778,000,000,  which  is  the 
number  of  foot  pounds  equivalent  to  1,000,000  B.T.U.  ;  this 
number  has  reached  a  value  of  0.23.  This  means  that  of  the 
heat  supplied  to  the  test  pumps  23  per  cent  is  utilized. 


TEST  OF  PUMPING  ENGINES  419 

The  preparation  of  a  pump  for  a  duty  test  consists  in  much 
adjustment  of  valve-gearing  and  other  steam  apparatus  which 
consumes  time.  Many  preliminary  runs  are  made  to  discover 
the  effect  of  changes  and  this  can  be  done  only  by  a  systematic 
method  of  keeping  data  and  making  alterations.  One  change 
should  be  made  at  a  time  and  its  effect  determined  before 
proceeding  to  another  change.  In  all  of  this  work  a  positive 
circulation  must  be  maintained  in  the  heating  coils  and  jackets, 
for  if  the  water  of  condensation  is  not  removed  these  cease  to 
operate  properly. 

To  illustrate  the  data  and  results  of  a  duty  test,  the  following 
has  been  taken  from  a  test  on  a  20,000,000  gallon  pump  at  the 
Lardner's  Point  Pumping  Station  in  Philadelphia,  conducted  by 
Francis  Head,  mechanical  engineer  of  the  Bureau  of  Water 
of  the  city  of  Philadelphia,  and  Edgar  G.  Hill,  mechanical 
engineer  representing  the  Holly  Manufacturing  Company. 

"At  Lardner's  Point  are  three  pumping  stations,  No.  i,  No. 
2  and  No.  3.  Station  No.  i  was  formerly  known  as  Frankford 
station,  and  contains  boilers  and  pumping  engines  that  have 
been  in  service  a  great  many  years,  which  are  now  retained 
only  as  a  reserve,  the  machinery  being  entirely  too  expensive 
to  keep  in  regular  service. 

"Stations  No.  2  and  No.  3  are  new  pumping  stations  through- 
out; the  buildings  and  equipment  therein  were  constructed 
for  the  new  nitration  distribution  system.  Each  consists  of  an 
engine  house  and  boiler  house.  The  two  stations  are  placed 
contiguous  and  for  all  practical  purposes  are  one  station. 

"In  engine  houses  No.  2  and  No.  3  are  twelve  vertical 
triple  expansion  self-contained  pumping  engines  of  20,000,000 
U.  S.  gallons  daily  capacity  each,  designed  for  a  normal  head  of 
225  feet,  but  capable  of  operating  economically  against  heads 
ranging  from  180  to  280  feet. 

"These  twelve  pumping  engines,  each  substantially  a 
duplicate  of  the  other,  designed,  constructed,  and  installed  by 
the  Holly  Manufacturing  Company  of  Buffalo,  N.  Y.,  make  an 
installation  the  largest  and  most  comprehensive  of  its  type,  not 
only  in  the  United  States  but  in  the  whole  world. 


420  PUMPING  MACHINERY 

"The  water  which  these  engines  pump  is  supplied  from  the 
nitration  plant  of  the  city  of  Philadelphia  at  Torresdale,  between 
two  and  three  miles  from  Lardner's  Point  pumping  station, 
through  an  underground  conduit  leading  from  the  Torresdale 
filtration  plant  to  the  Lardner's  Point  pumping  station.  It 
is  pumped  from  the  Lardner's  Point  pumping  station  through 
a  number  of  delivery  mains,  ranging  in  diameter  from  48  to  60 
inches,  to  various  parts  of  the  city  of  Philadelphia. 

"Ordinarly,  eight  or  nine  of  these  engines  discharge  the  water 
through  two  60  inch  delivery  mains,  which  are  so  connected 
together  as  to  form  substantially  one  main  of  twice  the  capacity 
of  a  60  inch  pipe,  against  a  head  of  approximately  185  feet.  The 
other  three  or  four  engines  ordinarily  pump  through  a  separate 
pipe  system  against  a  head  of  approximately  275  feet.  To 
show  the  action  of  the  governor  the  following  is  mentioned: 

"At  8  o'clock,  the  morning  of  December  10,  1909,  at  a  point 
about  one  thousand  feet  from  the  pumping  station,  a  12  foot 
length  of  the  60  inch  delivery  main,  into  which  eight  of  these 
pumping  engines  were  at  that  time  delivering  water  at  the  rate 
of  over  160,000,000  gallons  per  day,  against  a  head  of  approx- 
imately 185  feet,  split  from  end  to  end,  instantaneously  reducing 
•the  work  from  full  load  to  almost  no  load.  Five  of  these  pump- 
ing engines  were  in  Station  No.  2,  and  three  in  Station  No.  3. 

"All  of  the  pumping  engines  were  under  such  perfect  control 
of  the  governors  and  automatic  safety  devices  that  only  one 
of  the  eight  attained  sufficient  speed  to  make  the  automatic 
shutting-down  device  operative,  the  result  being  that  one 
engine  stopped  automatically,  and  the  other  seven  ran  at  a 
uniform  speed  until  gradually  closed  down  by  the  employees 
at  the  station.  Absolutely  no  damage  was  done  to  any  of  the 
engines  or  to  any  of  the  machinery  in  the  pumping  stations. 

"The  broken  60  inch  delivery  main  was  promptly  repaired 
and  the  engines  were  all  in  service  again  the  afternoon  of  the 
same  day  the  accident  occurred. 

"This  is  undoubtedly  the  first  time  the  governing  apparatus 
of  so  many  pumping  engines  has  been  subjected  simultaneously 
to  a  test  of  this  character.  That  none  failed  to  work  properly 


TEST  OF  PUMPING  ENGINES  421 

and  satisfactorily,  speaks  louder  than  words  for  the  reliability 
of  the  governing  apparatus. 

"One  of  the  twelve  engines  was  officially  tested  March 
9-10,  1910,  and  developed  a  new  high  duty  record  for  pumping 
engines  fitted  with  attached  jet  condenser,  on  a  24-hour  trial. 

"During  the  test,  readings  of  the  water  pressure  were  taken 
every  five  minutes  from  a  correct  pressure  gauge  checked  by  a 
mercury  column.  A  complete  round  of  observations  was  taken 
every  fifteen  minutes.  All  readings  and  weights  were  checked 
by  two  observers,  one  representing  the  city  and  one  the  company. 

"An  eight-hour  boiler  leakage  test  was  made  before  the  duty 
trial,  also  immediately  after  the  duty  trial. 

"A  summary  of  the  principal  results  is  given  below,  by  the 
makers: 

Size  of  engine 30",  60",  9o"x33"x66" 

Total  water  pressure,  Ibs 95-74 

Average  total  feed  water  per  hour,  Ibs 9,265.625 

Feed  water  corrections,  average  per  hour,  Ibs. : 

Drips  from  steam  header  and  steam  separator.  .  .  .        135-27 
Boiler  leakage,  average  of  two  tests 196.39 


Total.... ' -__.  33166 

Average  total. steam  per  hour  delivered  to  engine,  Ibs.  ...        8,933.07 

Entrainment,  per  cent •. i  .13  5 

Average  total  dry  steam  per  hour  delivered  to  engine 8,832.56 

Water  pumped  per  24  hours,  gallons 21,218,788 

Duty  per  1,000  Ibs.  steam  delivered  to  engine 182,382,200 

Duty  per  1,000  Ibs.  dry  steam 184,476,200  '* 

In  the  specifications  for  pumping  engines  definite  statements 
should  be  made  of  the  manner  of  conducting  the  test  and  of  the 
working  out  the  results.  As  an  example,  the  following  quota- 
tions are  made  of  clauses  from  the  specifications  for  the  Lard- 
ner's  Point  pump: 

"Section  119.  Duty  Test.  During  the  period  of  probation, 
and  before  its  final  acceptance,  each  engine  shall  be  subjected 
to  a  duty  test  of  twenty-four  (24)  hours'  duration.  The  test 
shall  be  conducted  by  two  (2)  engineers,  one  (i)  to  be  selected 
by  the  director,  and  the  other  by  the  contractor,  etc. 


422  PUMPING  MACHINERY 

"Section  120.  Determination  of  Head.  The  nead  for  com- 
putation of  duty  will  be  the  sum  of  the  head  in  feet  indicated 
on  a  gauge  attached  to  the  discharge  main  beyond  the  last  pump, 
and  the  vertical  elevation  of  the  center  of  this  gauge  above  the 
level  of  water  in  the  pump  well.  The  level  of  water  in  the  pump 
well  shall  be  determined  by  a  suitable  float  gauge,  and  no 
correction  or  allowance  will  be  made  for  any  friction  losses 
between  the  water  in  the  pump  well  and  in  the  discharge  pipe 
just  beyond  the  pump. 

"Section  121.  Total  Head.  The  total  head  for  the  contract 
duty  test  shall  be  not  less  than  two  hundred  and  fifteen  (215) 
and  not  more  than  two  hundred  and  twenty-five  (225)  feet. 

"Section  122.  Capacity.  The  capacity  of  the  engines  shall 
be  in  the  duty  test  not  less  than  twenty  million  (20,000,000) 
gallons  per  twenty-four  (24)  hours,  at  a  speed  of  not  more  than 
twenty  (20)  revolutions  per  minute.  The  capacity  of  the  pumps 
during  the  duty  tests  will  be  determined  by  plunger  displace- 
ment, and  no  correction  will  be  made  for  slip  unless  leakage 
from  plungers  and  valves  is  found  by  test  to  exceed  two  and 
one-half  (2^)  per  cent. 

"If  the  leakage  or  slip  is  found  by  Pitot  meter  to  be  more 
than  two  and  one-half  (2j)  per  cent,  the  capacity  and  duty 
shall  be  computed  from  the  measured  capacity. 

"Section  123.  Stuffing  Box  Leakage.  The  leakage  from  all 
the  water  stuffing  boxes  of  the  engine  shall  not  exceed  five 
hundred  (500)  gallons  per  hour.  This  shall  not  be  charged 
against  the  engine. 

"Section  124.  Feed  Pumps.  The  direct-connected  feed 
pumps  shall  be  operated  during  the  duty  trial,  supplying  water 
to  the  measuring  tanks,  and  running  against  the  usual  discharge 
pressure.  One  auxiliary  feed  pump  shall  be  operated  to  remove 
condensation.  No  allowance  will  be  made  for  these  pumps. 

"Section  125.  Steam  Measurement.  The  water  fed  to  the 
boilers  shall  be  weighed  and  any  condensation  in  the  lines  and 
drains  from  the  live  steam  separator  shall  be  deducted.  All 
steam  passing  through  the  separator  to  the  throttle  valve  shall 
be  charged  against  the  engine  as  dry  steam. 


TEST  OF  PUMPING  ENGINES  423 

"Section  37.  Uniformity  of  Steam  Diagrams.  The  con- 
struction and  adjustment  of  the  pump  valves  and  steam  valve 
gear,  and  the  balancing  of  the  plungers  and  other  moving  parts, 
shall  be  such  as  to  give  approximately  uniform  engine  indicator 
diagrams  from  each  of  the  steam  cylinders  on  the  up  and  down 
strokes." 

In  making  the  report  on  the  test  of  this  pump,  the  following 
statements  are  made: 

"The  level  of  water  in  the  pump  well  was  determined  by  a 
gauge  glass  placed  near  the  pump  suction  instead  of  a  float, 
as  this  gave  better  results. 

"Under  Section  122,  regarding  correction  for  slip,  as  before 
this  was  assumed  to  have  been  under  2-|-  per  cent,  and  no  cor- 
rection was  made  for  it. 

"Under  Section  123,  stuffing  .box  leakage  was  measured  and 
found  to  be  less  than  allowed. 

"Under  Section  124,  which  calls  for  the  operation  of  one 
auxiliary  feed  pump  to  remove  condensation,  the  results  of  the 
test  made  on  July  2oth,  1909,  reported  in  the  test  of  No.  16, 
were  used,  and  the  amount  so  determined  charged  against  the 
engine. 

"Under  Section  125,  the  same  six  inch  steam  line  was  used 
to  convey  steam  from  boiler  No.  26  to  the  pump.  As  before, 
the  drip  was  trapped,  condensed,  weighed  and  deducted  from 
the  feed  water. 

"The  correctness  of  the  feed  water  scales  was  checked  by 
Fairbanks  Standard  weights,  and  the  steam  and  water  gauges 
were  corrected  by  a  Crosby  dead  weight  tester,  which  had 
been  carefully  calibrated. 

"The  water  pressure  gauge  was  further  checked  by  a  mercury 
column  which  was  set  up  beside  the  pressure  gauge.  The 
density,  of  the  mercury  used  was  determined  in  the  testing 
laboratory  of  the  Department  of  Public  Works. 

"On  March  8th,  the  suction  and  discharge  valves  were 
tested  and  made  tight. 

"On  March  8th,  a  leakage  test  of  the  boiler  and  piping  was 
made  from  9.47  A.M.  to  5.47  P.M.  On  March  loth,  another 


424  PUMPING   MACHINERY 

test  was  made  from  i.oo  to  7.55  P.M.      The  average  net  loss 
was  196.392  pounds  per  hour. 

"The  capacity  of  the  engine,  as  actually  run  during  the  test, 
from  plunger  displacement,  was  21,218,481  gallons  per  24  hours 
at  a  speed  of  20.13125  R.P.M., 

"The  engine  worked  smoothly  and  satisfactorily  during 
the  entire  test,  and  the  requirements  of  the  specifications  as  to 
economy  and  capacity  on  the  duty  test  have  been  fully  complied 
with." 

The  following  data  was  obtained  during  test  from  average 
readings : 

DATA  AND  RESULTS 

Engine  tested,  contractor's  number 604 

Water  Bureau  number 14 

Date  of  test March  8,  9,  and  10,  1910 

Duration  of  test 24  hrs. 

CAPACITY 

Revolutions  during  test 28,989 

Average  revolutions  per  minute 20.13125 

Average  diameter  of  plungers,  inches % 32.985 

Average  stroke,  feet 5-496 1 

Number  of  plungers 3 

Displacement  per  revolution,  gallons 73 1.9494 

Displacement  per  24  hours,  gallons 21,218,481 

Displacement  per  24  hours,  at  contract  speed 21,080,144 

Water  used  to  lubricate  plungers,  per  hour,  gallons 360 

WORK  DONE 
Head  pumped  against — 

Pressure,  corrected  gauge 86.918  Ibs.  =200.50  ft. 

Suction  lift  to  center  of  pressure  gauge 8.822  Ibs.  =    20.35  ft. 

Total 95-740  Ibs.  =220.85  ft. 

Work  done  per  hour 1,629,397,423  ft.lbs. 


Duty 


DUTY 
Foot-lbs.  per  hr.  X  1000 


Net  steam  charged  to  pump  per  hour 
1,629,397,423,000 

8991.713 
=  181,211,013  ft.lbs. 


TEST  OF  PUMPING  ENGINES  425 

This    is    sEghtly   different    from    the    builders'    computations    given 

above. 

\ 

PRESSURES 

Throttle  gaug-e  reading,  Ibs.  per  sq.in 190.82 

Corrected  for  height 180. 19 

First  receiver,  Ibs.  per  sq.in 33.8 

Corrected  for  height 25.4 

Second  receiver,  inches  vacuum 9.5 

Vacuum  in  condenser  (reading  of  mercury  column)  inch 27.68 

Average  barometer  at  32°?.  sea-level 14-789  Ibs.  =30.06 

Average  barometer  at  floor  level  and  room  temperature  14.868  Ibs.  =30.20 


JACKET  AND  RECEIVER  DATA 

High-pressure  jacket,  Ibs.  per  sq.in 180.19 

Intermediate  jacket,  Ibs.  per  sq.in 34.5 

Low-pressure  jacket,  Ibs.  per  sq.in 0.5 

First  receiver  drip,  per  hour 422.16 

Second  receiver  drip,  per  hour 360.87 

Jacket  drip,  per  hour 660.62 

Total  drip,  per  hour 1443.65 

Per  cent  of  steam  passing  throttle -.-• 16.16 


TEMPERATURES 

In  exhaust  pipe,  4  ft.  below  L.P.  cylinder 113°  F. 

Water  pumped 40° 

Condensing  water '. 40° 

Water  leaving  condenser •  7° 

Water  after  passing  feed  heater  in  exhaust  pipe 85° 

Air  in  engine  room  at  mercury  column 80° 

Outside  average 3°° 


CALORIMETER 

Taken  after  test,  instrument  between  throttle  and  cylinder: 

Pressure,  Ibs J76-7 

Temperature 297.25°  F. 

Moisture,  per  cent I-I33 


426 


PUMPING  MACHINERY 


EVAPORATION 

Water  pumped  to  boiler  per  hour,  Ibs 9265.625 

Boiler  leakage  per  hour,  Ibs —  196.392 


Water  evaporated  per  hour 9069.233 

Engine  room  drip  per  hour —  135.270 


Steam  required  to  run  drip  pump  per  hour. 


+  57-75° 


Net  steam  charged  to  engine  per  hour 8991.713 

Boiler  horse-power  developed 324 

Boiler  rated  horse-power 500 


Head  End  L.P. 
10.07 


Crank  End  L.P. 
10.11 


To  accompany  Report  of  Daty  rrui  ut  11. 
Point  Pumping  Station  dated  March  21th  1910 
Cards  taken  10.30  P.M.  3-9-10 


H.P.  Water 
C2.C2 


Int.  Water 
61.00 


L.P.  Water 
63.40 


FIG.  337. — Indicator  Cards  from  Lardner's  Point. 


TEST  OF  PUMPING  ENGINES 


427 


"  The  boiler  used  was  an  Edgemoor  water  tube  boiler  No.  26 
of  the  same  size  and  type  as  used  on  the  previous  tests. 

' '  It  was  connected  to  the  engine  by  a  special  six  inch  steam 
pipe  which  was  well  covered. 

' '  Indicator  cards  are  arranged  to  correspond  to  the  position 
of  the  cylinders  from  which  they  were  taken.  These  are  shown 
in  Fig.  337. 

"The  distribution  of -power  between  the  different  cylinders 
as  determined  from  the  cards  is  shown  on  the  following  table: 


Steam. 

Head. 

Crank. 

Total. 

Per  cent  of 
Total. 

High-pressure  cylinder  .  . 
Intermediate  cylinder.  .  . 
Low-pressure  cylinder.  .  . 

163.47 
!35-35 
130.10 

168.55 
J33-75 

130.  12 

332.02 
269  .  10 

26O  .  22 

38.6 
3!-3 
31-1 

428  .  92 

432.42 

861.34 

100  .  0 

Mechanical  efficiency,    ratio   of    net   delivered  water  horse- 
power to  indicated  steam  horse-power 95-54 

Horse-power  from  volume  and  head  of  water 822  . 93 

Water  horse-power  indicated .  839  . 8 

Excess  of  indicated  over  delivered  water  H.P 2.1% 

Steam  passing  throttle  per  indicated   H.P.  per  hour,  Ibs...  10.37 

Steam  passing  throttle  per  delivered  water  H.P.  per  hour,  Ibs .  10 . 86 
Heat  units  from  steam  pressure  to  vacuum  temperature  used 

per  I. H.P.  per  minute 193  .04 

Heat  units  passing  throttle  from  steam  pressure  to  vacuum 

temperature  used  per  delivered  water  H.P.,  per  minute.  202.05 
Efficiency  from  heat  in  steam  delivered  to  engine  above  tem- 
perature in  exhaust  pipe  to  work  done  in  discharge  main .  20 . 99 

NOTE. — In  the  test  the  reheating  coils  in  both  receivers  were  out 
of  service." 

This  test  represents  one  of  the  highest  duties  obtained  on 
pumping  engines  and  should  be  studied,  as  many  results  ^an  be 
obtained  from  the  data. 


CHAPTER  X 
HIGH  DUTY  PUMPS    AND  WATER  WORKS  STATIONS 

To  study  the  modern  form  of  the  water  works  high  duty 
pump  and  its  installation  a  number  of  examples  will  be  given. 

The  Western  Pumping  Station  of  the  Water  Works  of  Cincin- 
nati, Ohio,  is  shown  in  the  Figs.  338  to  341.  The  station 
at  present  contains  three  25,000,000  gallon  pumps  for  lo;v 
service  and  three  12,000,000  gallon  pumps  for  high  service, 
built  by  the  Holly  Manufacturing  Co.  of  Buffalo,  N.  Y.  The 
station,  Fig.  339,  contains  space  for  an  additional  pump  of  each 
kind.  The  water  enters  through  a  gravity  tunnel  to  a  well,  and 
from  this  point  a  conduit  of  pipe  extends  to  the  suction  of  the 
various  units.  Two  suction  lines  are  taken  off  the  main 
conduit  at  each  pump,  one  of  them  passing  through  the  con- 
denser, and  these  enter  the  suction  valve  chambers,  as  will  be 
seen  later,  and  finally  connect  together  in  a  suction  air  chamber 
at  the  end.  The  two  discharges  from  each  side  of  the  pump 
are  connected  into  the  discharge  main,  which  is  connected  to  two 
service  mains  leaving  the  station  on  each  side  of  the  center. 

The  branches  from  the  sides  of  the  pump  may  be  cut  off 
by  valves  which  are  shown  by  the  conventional  cross,  and  by 
valves  in  the  service  lines;  any  of  these  may  be  cut  off  from 
the  station.  Valves  in  the  main  discharge  line  of  the  station 
make  it  possible  to  use  certain  of  the  pumps  on  one  service  main. 
Such  arrangements  are  valuable  at  times  for  service  and  for 
experimental  work.  Attention  is  called  to  the  fittings  left  for 
the  installation  of  the  new  pumps. 

The  steam  is  generated  in  the  boiler  room  in  which  the 
boilers  are  placed  on  the  side  away  from  the  pump  room.  This 
makes  a  longer  steam  line,  but  with  it  the  chimneys  can  be 
located  in  a  better  manner.  The  open  space  in  front  of  the 

428 


HIGH  DUTY   PUMPS  AND   WATER  WORKS  STATIONS     429 

boilers  is  so  wide  that  tubes  may  be  removed  from  the  boilers. 
This  space  forms  a  convenient  place  for  firing.    The  boilers  are 


FIG.  338. — Western  Pumping  Station 

connected  to  a  steam  header  carried  on  the  wall  between  the 
boiler  room  and  pump  room  and  branches  lead  to  the  engine. 


430 


PUMPING  MACHINERY 


FIG.  340. — Front 


of  Holly  Cincinnati  Pump. 

(To  face  page  431) 


HIGH   DUTY   PUMPS  AND  WATER   WORKS  STATIONS     431 

Each  branch  from  the  boiler  to  the  steam  main  has  two  valves, 
and  the  steam  main  has  a  number  of  valves  distributed  along  its 
length  so  that  certain  boilers  may  be  -used  with  a  given  pump  if 
desired.  The  branch  leading  to  each  pump  is  controlled  by  a 
valve  just  inside  the  engine  room  for  emergency  while  the 
main  stop  valve  is  placed  at  the  point  where  the  steam  enters 
the  high  pressure  cylinder.  The  former  valve  is  not  very  acces- 
sible, while  the  other  valve  used  in  starting  or  stopping  the 
pump  is  controlled  from  all  platforms  or  the  main  floor.  It  is 
better  to  place  the  high-pressure  cylinder  next  to  -the  boiler 
room  wall,  but  conditions  may  arise  so  that  this  is  not  advisable. 
With  the  water  suction  pipe"  in  the  position  shown  in  the  plan  of 
the  station  the  condenser  must  be  placed  on  this  end,  and  hence 
the  high-pressure  cylinder  is  placed  away  from  the  boiler  room 
and  the  reason  for  the  use  of  the  longer  pipe  is  evident. 

The  feed  water  for  the  boilers  is  taken  from  the  condenser 
and  then  pumped  through  economizers  by  the  pumps  shown  in 
the  boiler  house.  These  economizers  are  placed  behind  the 
boilers  in  the  flues  leading  to  the  stacks.  The  valves  controlling 
the  supply  of  water  to  these  and  the  method  of  discharging 
around  the  economizers  may  be  seen. 

If  forced  draft  is  required  to  make  up  the  draft  reduction 
by  the  economizers,  blowers  may  be  placed  in  the  boiler  room, 
forcing  air  beneath  the  boilers.  This  keeps  the  air  in  the  boiler 
room  in  circulation,  renewing  if  from  the  outside. 

The  small  electric  generators  placed  in  the  center  of  the 
building  are  used  for  lighting  and  for  power  for  the  cranes  and 
shops.  A  crane,  the  runway  of  which  is  seen  in  Fig.  338,  serves 
to  handle  the  parts  in  erection  or  repair. 

Fig.  339  shows  clearly  the  arrangement  of  offices,  machine 
shop,  store  room,  tool  room  and  toilet  rooms.  There  are  two 
rooms,  one  for  the  engineers  and  oilers  and  one  for  the  firemen. 
These  rooms  should  be  large,  airy  and  well  ventilated. 

A  study  of  this  plan  will  reveal  the  many  good  features  of 
the  design. 

The  pumps  shown  in  Figs.  340  and  341  for  this  station 
illustrate  a  type  of  high  duty  pump. 


432 


PUMPING   MACHINERY 


END  ELEVATION-LOW  SERVICE  25  MILLION  GALLON  ENGINE. 


FIG.  341. — Holly  Cincinnati  Pump. 

Water  enters  from  A,  passing  the  valves  D  on  each  side  of 
the  pump  and  entering  the  pipes  B,  which  are  continued  through 


HIGH  DUTY  PUMPS  AND   WATER   WORKS  STATIONS     433 


the  valve  chambers  HH  to  the  end  of  the  pump,  where  they 
combine  and  enter  the  suction  air  chamber  E.  The  water  is 
forced  out  by  the  plungers,  as  described  on  page  286,  and  is 
passed  through  the  pipe  /  which  is  made  a  portion  of  the  upper 
part  of  the  valve  chamber.  The  water  from  each  side  passes 
through  check  valves  F  and  finally  enters  the  discharge  pipe  at 
G.  The  discharge  check  valve  is  provided  with  a  small  by-pass 
valve  for  priming  the  pump. 

The  tops  of  the  valve  boxes  form  air  chambers  and  these 
are  supplied  with  compressed  air  when  necessary  through  the 
pipe  K.  The  steam  end  is  triple  expansion,  with  the  three 
cylinders  arranged  in  succession.  The 
cross  heads  are  equipped  with  four 
rods  extending  over  the  crank  and 
shaft  to  the  plunger  cross  head  below. 
The  end  cranks  are  overhung  but 
carry  return  cranks  which  are  con- 
nected to  the  end  cranks  on  the 
shaft  M.  These  return  cranks  are 
so  placed  on  the  pins  that  they  turn 
the  shaft  M  positively,  that  is,  when 
the  cranks  on  one  end  are  passing  a 
dead  point  for  the  connecting  rod,  the 
others  are  at  maximum  throw.  The 
shaft  M  is  used  to  operate  eccentrics, 
the  motions  of  which  serve  to  oper- 
ate the  Corliss  valves  of  the  cylinders. 

In  high  duty  engines  the  cylin- 
ders are  usually  jacketed  and  in 
many  cases  the  receivers  between  the 
cylinders  are  heated  by  coils.  Fig. 
342  shows  the  internal  arrangement 
of  a  reheating  receiver.  The  coil  is 

drained  by  means  of  a  trap  and  this 

.   .      ,  FIG.  342. — Reheater  Receiver. 

must  act  positively. 

Fig.  343  is  the  interior  of  the  Central  Park  Avenue  Pumping 
Station  in  Chicago,  which  is  similar  to  the  Springfield  Avenue 


434 


PUMPING  MACHINERY 


FIG.  343. — Central  Park  Avenue  Station. 


FIG.  345. — Worthington  Pumps  at  Fall  River,  Mass. 


N     6'Steam  Pipe  to  Frest pt  Header 


42  Delivery  Elev.+20.5J| 


6  Steam  Pipe  to  Present  Engine   1202 


FIG.  344. — Word 


Gallery  Elev.+27V 


ijrh  Water  Kiev. +  5 


Bottom  of  Pit  Elev.-18'O" 


n  High  Duty  Pump. 


(To  face  page  435) 


HIGH    DUTY  PUMPS  AND   WATER    WORKS  STATIONS     435 

station.  The  pumps  at  the  latter,  shown  in  Fig.  344,  developed 
a  duty  of  174,735,801  foot  pounds  per  1,000  pounds  of  steam. 
Water  is  brought  to  this  station  through  a  long  tunnel  from 
Lake  Michigan  and  discharges  into  a  well  built  in  the  room. 
The  suction  pipes  are  provided  with  foot  valves  and  the  surface 


FIG.  346. — Allis  Pumps  at  the  Baden  Station. 

condenser  uses  water  which  is  taken  into  the  suction  for  condens- 
ing water.  The  pump,  as  described  earlier,  is  a  duplex,  triple 
expansion,  direct-acting  pump,  supplied  with  compensators. 
The  cylinders  are  all  in  tandem  and  the  weight  is  carried  by  a 
balancing  piston  so  connected  to  the  discharge  main,  as  shown  in 
Fig.  344,  that  should  a  break  occur  in  the  main  the  pump  will 
come  to  rest.  These  pumps  may  be  placed  in  a  small  floor  area. 


436  PUMPING  MACHINERY 

Fig.  345  shows  a  station  equipped  with  a  Worthington 
horizontal,  triple-expansion,  direct-acting  engine.  This  pump 
was  described  on  page  100. 


FIG.  347. — Allis  Pumps  at  Bissel's  Point. 

Fig.  346  shows  the  appearance  of  the  Allis-Chalmers  pump, 
installed  at  the  Baden  station  of  the  St.  Louis  water  works, 
while  Fig.  347  shows  their  pump  at  the  Bissel's  Point  station 


HIGH  DUTY   PUMPS  AND  WATER   WORKS  STATIONS     437 

of  th-j  same  system.  In  Fig.  347  the  return  crank  driving  the 
valve  shaft  on  the  first  balcony  is  clearly  seen.  The  valves  of 
both  these  pumps  are  of  the  Corliss  type  on  the  high-pressure 


FIG.  348. — Holly  Pump  at  Boston. 

and  intermediate-pressure  cylinders,  but  on  the  low-pressure 
cylinder  poppet  valves  are  used. 

Fig.  348  shows  a  pump  built  by  the  Holly  Company  for  the 
Spot  Pond  Station  of  the  Boston  water  works.  In  this  pump 
the  valve  or  eccentric  shaft  is  driven  by  bevel  gears.  The  exhaust 
valve  of  the  intermediate-pressure  cylinder  of  this  pump  and 


438 


P  UMPI NG    M  A  CHI  NER  Y 


both  valves  of  the  low-pressure  cylinder  are  of  the  poppet  type. 
The  figure  shows  the  various  gauges  and  valves  as  well  as  the 
governor  which  actuates  the  cut-off  on  the  high-pressure  and 


FIG.  349. — Allis-Chalmers  Pump. 

intermediate-pressure  cylinders.    In  all  these  pumps  the  bracing 
of  the  frame  is  to  be  noted. 

Fig.  346  shows  clearly  the  type  of  complete  frame  often 
used  with  these  pumps  and  in  them  the  valve  chambers  are 
separate  from  the  frame.  In  Fig.  349,  showing  an  Allis-Chalmers 
pump,  the  valve  chambers  are  used  for  supports.  In  the  Holly 


HIGH  DUTY   PUMPS  AND   WATER  WORKS  STATIONS     439 

pump  of  Fig.  340,  the  separate  air  chambers  are  used,  the  bed 
plate  at  the  floor  line  being  carried  by  the  inclined  frames. 
A  small  2,500,000  gallon  pump  built  by  the  Holly  Manu- 


FIG.  350. — Holly  Pump  at  Washington,  D.  C. 

facturing  Company  for  the  Trumbull  Street  Pumping  Station  of 
Washington,  D.  C.,  is  shown  in  Fig.  350.  This  pump  gave  a  duty 
of  164,644,000  foot-pounds  per  1,000  pounds  of  dry  steam. 


440 


PUMPING  MACHINERY 


The  valve  shaft  in  this  engine  is  oscillated  by  an  eccentric  on  the 
main  shaft.    A  governor  controls  the  cut-off  on  the  high-pressure 


HIGH   DUTY  PUMPS  AND   WATER    WORKS  STATIONS     441 


cylinder,  but  the  cut  off  on  the  other  cylinders  is  controlled  by 
small  hand  wheels  seen  beneath  the  upper  platform.  All  of  the 
valves  are  of  the  Corliss  type. 

Fig.  351  represents  the  arrangement  of  a  station  in 
Brooklyn,  N.  Y.  The  boiler 
houses  are  placed  at  the  ends  of 
the  pump  house  and  the  steam  is 
taken  to  the  engines  by  the  main 
steam  supply.  The  water  ends 
are  connected  to  a  suction  main 
on  one  side  of  the  building  and 
discharges  are  carried  from  each 
pump  separately.  In  the  rooms 
on  the  side  of  the  pump  house, 
as  indicated  by  outline  only,  are 
placed  the  offices,  shops,  toilets 
and  lockers,  while  the  houses 
beside  the  boiler  rooms  are 
intended  for  coal  storage. 

The  storage  of  coal  is  import- 
ant in  all  plants  for  the  supply 
of  water  or  for  any  other  pub- 
lic service.  This  matter  cannot 
receive  too  much  attention.  A 
supply  for  a  week's  consumption 
should  be  provided  for  and  if 
conditions  of  coal  delivery  are 
poor  this  should  be  increased. 
In  some  cases  a  month's  supply 
is  none  too  much.  The  coal 
should  be  stored  beneath  cover 
as  the  alternate  wetting  and  dry- 
ing when  uncovered  causes  the 
coal  to  deteriorate.  With  soft 
coal  the  pile  should  be  ventilated.  FlG'  35  2--MemPhis  Station- 

The  arrangement  of  the  station  with  the  boiler  room  at  one 
end  is  not  advisable  if  conditions  are  such  that  the  arrangement 


442  PUMPING  MACHINERY 

shown  in  Fig.  339  is  possible.  When  possible  an  arrangement 
with  the  pump  room  and  boiler  room  beside  each  other  will 
make  a  better  plan  for  economy  and  for  ease  of  superintending. 
In  this  case  the  plant  is  more  compact  and  in  a  short  time  the 
chief  engineer  may  have  a  complete  view  of  the  whole  plant. 

Fig.  352  shows  a  plant  at  Memphis,  Tennessee,  where  the  first 
vertical  Worthington  high  duty  pumps  were  installed.  The 
plant  was  limited  in  ground  area  as  a  given  capacity  had  to  be 
placed  in  this  space.  The  vertical  high  duty  pump  installed 
was  of  the  compound  duplex  type,  and  on  test  in  1891,  this  pump 
gave  117,325,000  foot-pounds  per  1,000  pounds  of  steam.  The 
design  was  well  thought  out  and  as  shown  four  10,000,000  gallon 
pumps  were  placed  in  a  pit  38  feet  in  diameter.  The  steam 
cylinders  were  30  and  60  by  48  inches,  while  the  water  cylinders 
were  28  by  48  inches.  The  suction  well  is  shown  in  the  figure. 
The  pit  is  founded  on  a  heavy  concrete  base  and  the  wall  is 
built  circular  so  that  it  need  not  be  as  thick  as  an  ordinary 
letaining  wall.  % 

Fig.  353  shows  the  arrangement  of  the  water  works  at 
Zurich,  Switzerland,  as  described  in  the  Engineering  News,  Vol. 
32,  page  34.  This  plant  is  operated  by  turbines,  the  vertical 
shafts  of  which  drive  the  bevel  wheels  E,  which  arc  placed  on 
short  counter-shafts.  These  counter-shafts  drive  the  main  shaft 
B  by  means  of  the  gears  D,  and  the  gears  C  drive  the  shafts  of 
the  pump  which  are  blocked  in  at  A  A.  The  pumps  are  of  the 
two-crank  type,  a  crank  being  at  each  side  of  the  pump  with  the 
pins  quartering.  These  pumps  are  connected  to  four  different 
services  so  that  any  pump  can  be  used  for  any  pressure.  Of 
course  this  means  that  the  pump  power  will  change  and  this 
fact  explains  the  reason  for  the  long  jack-shaft.  A  pair  of  pumps 
which  would  be  operated  by  a  pair  of  turbines  properly  at  one 
pressure  would  not  operate  if  it  were  desired  to  use  a  higher 
pressure.  When  such  is  the  case,  the  additional  power  is  obtained 
from  other  turbines.  Although  the  use  of  the  jack-shaft  means 
a  certain  amount  of  friction,  the  flexibility  of  the  station  is 
increased  so  much  by  it  that  the  installation  is  proper. 

A  small  auxiliary  boiler  and  engine  may  be  used  when  the 


HIGH   DUTY  PUMPS   AND    WATER   WORKS  STATIONS     443 


water  wheels  are  out  of  commission.    These  are  shown  at  H  and 
G,  and  at  WW  are  high  pressure  water  wheels  which  may  be 


444 


PUMPING  MACHINERY 


run  from  the  high-pressure  reservoir  when  needed  to  operate 
electric  generators  00  for  lighting  at  night.  The  station  also  is 
connected  to  a  rope  transmission  at  R. 

Fig.  354  shows  the  arrangement  of  air  compressors  for  the 
Harris  pump  described  in  Chapter  XIII.     The  steam  cylinder 


FIG.  354, — Harris  Air  Pump  Station  at  High  Bridge. 

A  receives  steam  from  a  boiler  house  not  shown,  while  the  air 
cylinder  B  discharges  its  air  into  one  of  the  pipes  CC  and  draws 
its  supply  from  the  other.  The  action  of  the  pump  is  described 
later,  but  the  plant  is  shown  at  this  point  so  that  the  arrange- 
ment of  machinery  may  be  seen. 

The  station  shown  in  Figs.  355  and  356  is  located  in  Mil- 


HIGH  DUTY  PUMPS  AND   WATER  WORKS  STATIONS     445 

waukee,  Wis.,  and  is  used  for  flushing  the  Milwaukee  river. 
In  the  tower  room  are  the  "cylinders  of  the  steam  engine,  which 
is  used  to  drive  a  screw  pump  which  lifts  30,000  cubic  feet  of 
water  (224,000  gals.)  per  minute  against  a  head  of  3^  feet.  The 
water  is  drawn  through  a  tunnel  from  Lake  Michigan  and 
forced  into  Kinnickinnic  River.  The  tunnel  passes  under  the 
power  house  and  turns  after  passing  the  screw  of  the  pump. 


FIG.  355. — Kinnickinnic  River  Station. 

The  boiler  house  is  equipped  with  two  return  tubular  boilers 
furnished  with  superheaters.  The  engine  is  a  tandem  compound 
engine  and  the  screw  which  was  shown  in  Fig.  107  is  I2-J-  feet  in 
diameter. 

The  boiler  room  and  coal  room  are  seen  in  the  plan  together 
with  toilet  room  and  locker  room. 

The  Lardner's  Point  Pumping  Station  of  the  city  of  Phila- 
delphia consists  of  three  separate  pumping  stations,  Figs.  357, 
358  and  359.  "  The  first  is  an  old  station  formerly  termed  the 
"  Frankford  Pumping  Station  "  and  was  used  in  the  old  system 
to  pump  water  from  the  Delaware  River  to  the  Frankford  dis- 


446 


PUMPING   MACHINERY 


tribution  system.  It  is  now  termed  No.  i  House,  and  the  con- 
nection with  the  river  has  been  closed  and  a  new  connection 
made  to  the  filtered  water  conduit  leading  from  the  outlet 
shaft  of  the  Torresdale  conduit  from  the  filter  beds. 

"  No.  i  pumping  station,  or  the  old  '  Frankford,'  consists 
^f  one  compound  vertical  Cramp  pump  of  ten  million  gallons 
capacity,  one  Wetherill  horizontal,  ten  million  gallons  capacity, 


FIG.  356. — Kinnickinnic  River  Station. 

one  Southwark  vertical,  twenty  million  gallons  capacity  and  one 
Southwark  vertical  horizontal,  fifteen  million  gallons  capacity. 
For  this  station  there  are  twelve  marine  type  boilers  of  200  H.P. 
capacity  each. 

' '  Two  entirely  new  stations  were  constructed  and  contain 
twelve  (12)  vertical  triple  expansion  Holly  pumping  engines  of 
twenty  million  gallons  daily  capacity  each.  The  engine  rooms 
are  built  separate  from  the  boiler  houses  and  are  171  feet  long  by 
87  feet  wide,  constructed  of  gray  standard  size  brick  trimmed 


HIGH    DUTY   PUMPS    AND  WATER   WORKS  STATIONS     447 


448 


PUMPING   MACHINERY 


FIG.  358. — Lardner's  Point  Station. 


FIG.  359. — Lardner's  Point. 

with  granite  and  terra  cotta.    All  the  roof  coverings  are  of  red 
tile.    The  water  ends  of  the  pumps  are  set  in  the  basement  under 


HIGH  DUTY   PUMPS  AND   WATER  WORKS  STATIONS     449 

the  floor  of  the  engine  room,  and  the  entire  steam  ends  are  all 
above  the  floor  level.  The  pump  well  is  located  under  the 
basement  floor  in  the  center  of  the  engine  houses,  extending 
their  full  length.  It  is  constructed  of  reinforced  concrete,  horse- 
shoe shaped  in  section,  14  feet  in  width  and  height. 

"Between  engine  houses  Nos.  2  and  3  a  gate  chamber  is 
located  which  controls  the  discharge  from  the  larger  connection 
to  the  outlet  shaft  of  the  Torresdale  conduit.  It  is  connected 
to  the  pump  well  of  both  houses,  and  gates  have  been  installed 
for  connecting  the  pump  well  of  a  future  house  to  be  located 
west  of  the  present  plant. 

* '  The  boiler  houses  of  the  new  stations  are  of  the  same  general 
architecture  and  contain  the  following: 

House  No.  2\ 

Six    Edgemoor   water    tube    boilers,     500   horse-power 

capacity  each,  equipped  w.ith  Wetzel  stokers. 
Twelve  Internal-fired  tubular  boilers,   200  horse-power 
capacity  each. 

House  No.  j: 

Eight  Edgemoor  boilers,  500  horse-power  capacity  each, 
equipped  with  Wetzel  stokers  and  two  Green  econ- 
omizers. 

"  In  the  annex  of  boiler  house  No.  2  are  three  50  K.W. 
generators  each,  furnishing  light  for  the  entire  station  and 
current  for  the  coal  handling  machinery,  electric  crane,  etc. 

' '  The  smoke  flues  of  the  boiler  houses  are  connected  to  four 
brick  chimneys  150  feet  high  and  7  feet  internal  diameter — two 
for  No.  2  and  No.  3  houses  of  the  Custodis  and  Heinicke  type 
respectively. 

"The  coal  is  delivered  to  overhead  pockets  of  3,000  tons 
capacity  in  the  boiler  houses  of  Nos.  2  and  3  stations  by  means 
of  a  tower  and  belt  conveyor  with  capacity  for  handling  50  tons 
per  hour.  Coal  may  be  received  either  by  rail  or  by  boat,  and 
the  general  design  of  this  equipment  is  similar  to  the  one  at 
Torresdale. 

"Data  relating  to  engines  and  boilers  in  Nos,  2  and  3  houses, 
Lardner's  Point  pumping  station; 


450  PUMPING  MACHINERY 


ENGINES 

Nominal  capacity  of  each  engine 20,000,000  gallons  daily. 

Number  of  revolutions  per  minute 20 

{Stroke si  ft. 

Piston  speed,  feet  per  minute 220 

High.  Intermediate          Low. 

Cylinder  diameter 32"  60"  go" 

Diameter  piston  rod i\"  7$"  7$" 

First.  Second. 

Receiver  volume 205  cu.ft.         304  cu.ft. 

Receiver  heating  surface 166  sq.ft.         304  sq.ft. 

Diameter.  Length. 

Cross-head  pins 12"  n" 

Crank  pins 12"  n" 

Shaft  bearings 17$"  32" 

Shaft  at  center 2o£" 

Distance  rods. — Four  (4)  each,  5  inches  diameter. 

Air  pump. — One  (i)  28-inch  diameter,  66  inches  stroke. 

Feed  pump. — One  (i)  3^-inch  diameter,  66  inches  stroke. 

Feed  water  heater. — One  (i)  in  exhaust,  308  sq.ft. 

Flywheels. — Two    (2-)    20    feet    diameter    and    weighing    32    tons    each 

(approximate) . 

Throttle  valve. — 8  inches  diameter. 
Exhaust  pipe. — 24!  inches  diameter. 

Suction  pipe. — Main  42  inches  diameter,  branch  30  inches  diameter. 
Discharge  pipe. — Main  42  inches  diameter,  branch  30  inches  diameter. 
Suction  injection. — 8  inches  and  10  inches  diameter. 
Force  injection. — 3  inches  and  3^  inches  diameter. 
Overflow. — 18  inches  diameter. 
Diameter  of  plungers. — 33  inches. 

No.  2  House.    No.  3  House 
Number  of  pump  valves 960  864 

BOILERS 
Furnace  Flue  Tubular: 

Number  of  boilers 12 

Diameter  of  shell 108  ins. 

Length  of  shell 20  ft. 

Thickness  of  shell If  in. 

Number  of  fire  boxes .       2 

Diameter  in  inches 41 

Length  of  fire  boxes 8  ft. 

Number  of  tubes 195 

Length  of  tubes 9  ft.  3  ins. 

Diameter  of  tubes 3  \  ins. 

Length  of  grate 5  ft.  5!  ins. 

Area  of  grates 40 J  sq.ft. 


HIGH  DUTY  PUMPS  AND   WATER  WORKS  STATIONS     451 

Water  Tube  Boilers: 

Number  of  boilers 14 

Stoker Wetzel. 

Number  of  tubes 252 

Length  of  tubes 18  ft. 

Diameter  of  tubes 4  ins. 

Steam  drums  (two) 36  ins.  diam. 

Length  of  steam  drum 2 1  ft. 

Length  of  grate *  8$  ft. 

Area  of  grate 102  sq.ft. 

Fig.  360  is  an  exterior  view  of  the  Queen  Lane  station  of 
the  Philadelphia  water  works  and  is  shown  to-  illustrate  the 
appearance  of  a  station. 

Fig.  361  shows  a  Riedler  water  works   pump  used  in  the 


FIG.  360. — Queen  Lane  Station. 

water  works  of  the  city  of  Berlin,  Germany.  The  large  positively 
closed  Riedler  valves  are  seen  and  the  tubes  beneath  the  suction 
valves  prevent  the  suction  from  breaking  and  make  the  upper 
part  of  the  suction  chamber  an  air  chamber.  The  plunger  is 
passed  through  a  long  sleeve  and  although  different  from  the 
Worthington  plunger  and  ring,  it  is  the  equivalent  of  it.  The 
large  air  chambers  on  the  discharge  reduce  the  variation  in 
water  pressure. 


452 


PUMPING  MACHINERY 


HIGH  DUTY   PUMPS  AND   WATER   WORKS  STATIONS     453 


The  steam  cylinders  show  the  poppet  valve  form  of  cylinder 
so  -extensively  used  in  Europe. 

One  of  the  Jersey  City  water  works  stations  is  equipped 
with  this  type  of  pump.  The  station  is  shown  in  Fig.  362. 

The  pumps  are  driven  by  turbines  or  steam  cylinders  placed 
in  tandem  with  the  water  ends.  There  are  three  units  installed 
in  the  station,  two  being  high-pressure  pumps  and  one  a  low- 
pressure  pump.  The  suction  is  taken  from  the  fore  bay  by 
pipes  on  each  side  of  the  unit  and  the  two  discharge  air  chambers 
are  connected  to  the  discharge  main,  which  is  carried  to  the 
end  of  the  building. 


Name. 

Station. 

Date. 

Duty  per  1,000 
Ibs.  Dry  Steam. 

Savery  

5,000,000 

Smsaton  Newcomen  .  . 
Watt               



— 

I2,OOO,OOO 
20,000,000 

Cornish  engine  
Simpson's 



— 

60,000,000 
8^,000,000 

Holly  Quadruplex  .  ,  . 
Gaskill 



— 

85,000,000 
1  1  7,936,698 

Worthington 

120,000,000 

Reynolds    

1  18,327,041 

Reynolds-Hannibal.  . 
Reynolds  
Reynolds  
Worthington  (Direct) 
W7orthington  (Direct) 
Allis-Chalrners           .  . 

North  Point. 
St.  Louis. 
Chicago. 
Chicago. 
St.  Louis. 

Nov.    3-4,    1903 
Aug.  27-28,  1902 

154,048,700 
154,048,704 
179,454,250 
161,676,942 
I74,734,8oi* 
181,000,000 

Holly 

Washington,  D.C. 

Nov.     2-3,  19015 

162,644,000 

Holly 

Cleveland,  Ohio. 

Dec.          5,  1907 

164,642,226 

Holly                 

Brockton,  Mass. 

Oct.  14-15,  1909 

170,000,000 

Holly 

Louisville,  Ky. 

May      1-2,  1909 

195,020,000* 

Holly 

Albany,  N.  Y. 

May  29-30,  1909 

182,281,000 

Snow               

Mahanoy  City,  Pa. 

Nov.       12,  1906 

141,369,439 

Allis-Chalmers  
Allis-Chalmers  
Holly             

St.  Louis. 
Boston,  Mass. 
Philadelphia,  Pa. 

Feb.        26,  1900 
May      1-2,  1900 
Mar.          9,  1910 

179,454,255 
178,497,000 
184,476,200 

Worthington  (Direct) 
Worthington  (Direct) 

Fall  River. 
Montreal. 

Sept.  23,        1909 
Nov.  27,        1909 

136,500,000 
177,538,000* 

*Superheat 


The  12  inch  steam  pipe  and  the  24  inch  exhaust  pipe  are 
carried  in  a  trench  on  one  side  of  the  station  and  the  exhaust 
from  all  the  engines  is  cared  for  by  one  large  jet  condenser. 


454 


PUMPING  MACHINERY 


HIGH  DUTY  PUMPS  AND   WATER  WORKS   STATIONS    455 

In  such  a  plant  the  engines  are  only  used  in  case  of  low  water,  and 
at  other  times  the  pump  rods  are  disconnected  from  the  steam 
piston  rod  so  that  the  piston  does  not  act  as  a  resistance  to 
motion. 

To  give  some  idea  of  the  value  of  the  duty  of  modern  pump- 
ing engines,  see  table  on  page  453. 


CHAPTER  XI 
SPECIAL  PUMPING  MACHINERY 

THERE  are  several  special  pumps  used  to  handle  liquids 
which  should  be  discussed.  These,  although  the  same  in  principle 
as  those  which  have  been  described,  have  particular  features 
which  make  them  noteworthy. 

Condenser  Pumps  are  of  two  kinds:  Circulating  pumps  and 
air  pumps.  The  circulating  pump  is  used  to  force  the  condensing 
water  through  the  condenser,  while  the  air  pump  is  used  to 
remove  the  condensed  steam  and  air  from  the  steam  side  of  a 
condenser,  or,  in  the  case  of  a  dry  air  pump,  the  pump  handles 
air  and  water  vapor  only,  the  condensed  steam  being  removed 
from  the  condenser  by  a  wet  pump  or  by  gravity  as  in  a  barome- 
tric condenser.  The  circulating  pump  in  many  cases  lifts  the 
water  a  short  distance  only,  and  for  that  reason  it  is  built  as  a 
tank  pump,  while  in  other  cases  it  may  be  used  to  overcome 
the  friction  of  a  long  pipe  line  to  and  from  the  cooling  tower 
as  well  as  the  lift  to  the  top  of  the  tower,  in  which  case  it  must 
be  heavier. 

The  circulating  pump  most  often  takes  the  form  of  a  centrif- 
ugal pump,  as  it  has  to  lift  large  quantities  of  water,  and  space 
is  limited. 

Fig.  363  shows  a  combined  air  and  circulating  pump  of  the- 
Wheeler  Condenser  and  Engineering  Company.     In  this  pump 
the  two  water  cylinders  are  placed  in  tandem  with  the  steam 
cylinder  which  takes  the  form  of  any  of  the  simplex  pumps. 

In  Fig.  363  the  Kno\vles  pump  is  shown.  The  circulating 
pump  forms  one  end  support  for  the  condenser.  The  water  is 
discharged  through  A  into  one  set  of  tubes  and  then  it  retunis 
through  B  and  the  upper  set  of  tubes  to  C,  where  it  discharges. 
The  air  pump  forms  the  other  support  for  the  shell.  It  takes  the 
air  and  water  from  the  condenser  and  discharges  it  through 
/),  The  suction  space  F  is  connected  to  G. 

456 


SPECIAL  PUMPING  MACHINERY 


457 


To  find  the  size  of  the  water  and  air  ends  of  the  pump,  suppose 
that  W  pounds  of  steam  per  hour  at  a  pressure  p  are  to  be 
condensed.  If  r  is  heat  of  vaporization  of  the  steam,  x  its 
quality,  4°  the  temperature  of  the  condensed  steam,  and  q 
the  heat  of  the  liquid,  and  if  G  pounds  of  water  entering  at 


FIG.  363. — Combined  Air  and  Circulating  Pump. 

and  leaving  at  tQ°  F  are  to  be  used,  G  is  given  by  the  equation: 


qt0-qa 

If  the  number  of  revolutions  or  double  strokes  N  are  assumed, 
the  displacement  of  the  water  end  will  be 

G 


The  air  end  of  the  pump  is  made  in  many  cases  of  empirical 
design.  Some  authors  give  ratios  of  volume  displaced  by  the 
pump  per  minute  to  the  volume  of  the  condensed  steam  or  to 
the  volume  of  the  low-pressure  cylinder  of  the  engine  which  is 
discharging  into  the  condenser.  Several  of  these  are  mentioned. 


458  PUMPING   MACHINERY 

RATIO  OF  Am  CYLINDER  DISPLACEMENT  TO  LOW-PRESSURE 

CYLINDER. 

Single-acting  vertical  pump  surface  condenser i  9 

Single-acting  vertical  pump  jet  condenser i  13 

Double-acting  horizontal  pump  surface  condenser.  ...    i  15 

Double-acting  horizontal  pump  jet  condenser i  12 

Double-acting  horizontal  pump-compound  engine  sur- 
face condenser i  26 

Single-acting  horizontal  pump-compound  engine  sur- 
face condenser i  16 

RATING  OF  AIR  CYLINDER  DISPLACEMENT  TO  VOLUME  OF 
CONDENSED  STEAM. 

Surface  condenser 20  :  i 

Jet  condenser 40  :  i 

This  may  be  a  satisfactory  way,  but  it  is  better  to  estimate 
the  volume  from  the  air  probably  present.  Water  usually  con- 
tains air  to  about  one-fifteenth  of  its  volume.  This  amount  of 
air  is  at  atmospheric  pressure  pa  and  it  must  be  cared  for  by 
the  air  pump  at  a  reduced  pressure.  In  addition  to  this  there 
are  small  leaks  in  the  pipe  line  which  allow  more  air  to  enter.  A 
small  hole  will  destroy  the  proper  acting  of  the  air  pump.  The 
author  at  one  time  desired  to  run  a  2000  H.P.  condensing  engine 
with  no  vacuum  and  still  use  the  condenser  to  operate  so  as  to 
measure  the  steam  consumption,  and  he  found  that  a  2-inch  plug 
removed  from  a  30  inch  exhaust  main  was  sufficient  to  destroy 
the  vacuum  completely  with  the  air  pumps  running  at  full 
speed.  To  find  the  volume  of  air  per  minute  the  following 
formula  will  be  used,  allowing  100  per  cent  for  leakage. 


p=  absolute  pressure  in  the  condenser,  Ibs.  per  sq.in. 
ps=  vapor  tension  or  absolute  steam  pressure  correspond- 
ing to  Tc 

Tc  =  absolute  temperature  in  condenser 
Ta = absolute  temperature  atmosphere. 
W  =  weight  of  steam  per  hr.  or  weight    of   water   and 
steam  per  hr.  in  a  jet  condenser. 


SPECIAL  PUMPING  MACHINERY 


459 


This  equation  shows  the  importance  of  making  pt  as  different 
from  p  as  possible.  The  terms  p  and  ps  do  not  differ  much,  and 
by  taking  the  mixture  of  air  and  vapor  on  its  way  to  the  air 
pump,  through  as  cold  a  passage  as  possible,  the  term  pa  is 
made  smaller  and  the  denominator  is  increased,  making  V  small. 
This  is  the  reason  for  the  great  advantage  in  counter  current 
for  condensers,  and  even  in  the  condenser,  shown  in  Fig.  363, 
the  coldest  water  should  enter  directly  over  the  air  pump  inlet 
so  as  to  cool  the  mixture  going  to  the  pump. 

From  the  volume  thus  com- 
puted the  displacement  of  the 
air  pump  is  given  by: 


Knowing  the  displacements 
of  these  pumps  a  stroke  may  be 
assumed,  and  from  it  the  area 
determined. 


I1 


Aap  — 


D, 


ap 


FIG.  364. — Cards  from  Air  Pump  and 
Circulating  Pump. 


The  cards  from  the  water  end 
are  shown  in  the  lower  part  of 
Fig.  364,  while  those  for  the  air  end  are  shown  above.  The 
combination  of  these  or  the  addition  of  them  when  reduced  to 
the  proper  scale,  on  account  of  the  difference  in  piston  area, 
will  give-  the  total  work,  and  from  this  the  size  of  the  steam 
cylinder  is  given,  if  the  M.E.P.  be  found  for  a  given  boiler 
pressure.  Allowing  33  per  cent  for  friction,  which  is  made  large 
to  give  a  certain  driving  power,  the  following  results: 


(M.E.P. 


(M.E.P.  )PAP 


(i.oo-o.33)(M.E.P.)8 

Separate  air  pumps  are  often  used.    Fig.  365  shows  a  steam 
driven  pump   used   in   the   U.  S.    Navy.     This  air  pump   is 


460 


PUMPING  MACHINERY 


FIG.  365. — Air  Pump. 


SPECIAL  PUMPING  MACHINERY  461 

made  with  two  air  cylinders  driven  through  gears  from  a 
steam  cylinder  placed  on  one  side  of  a  pump  barrel.  The 
pump  is  of  the  bucket  type  with  foot  valves  A  A  and  head 
valves  at  B.  These  with  the  valves  in  the  bucket  at  C  are 
all  spring-controlled  metal  valves.  The  foot  valves  are 
placed  on  an  inclined  partition  for  the  purpose  of  making  it 
easier  to  discharge  the  air  when  the  piston  rises  and  forms 
a  vacuum.  The  lip  around  the  discharge  valve  matyes  a  dam 
and  covers  the  valve  with  water.  This  makes  thenr  air  tight. 
The  other  valves  are  in  that  condition,  sirice  all  of  the  water 
on  the  bucket  or  that  over  the  foot  valves  can  not  be  driven 
out,  as  the  valves  limit  the  motion  of  the  bucket.  On  the  down 
stroke  of  the  bucket  the  pressure  in  the  space  above  it  soon 
falls  to  a  low  vacuum  because  it  had  been  completely  filled 
with  water;  this,  then,  causes  the  valves  to  open  and  take  air 
from  the  lower  portion  of  the  cylinder.  The  air  in  the  water 
also  separates  and  rises  to  the  top  of  the  cylinder.  Finally 
trie  bucket  reaches  the  water  below,  and  this  is  driven  through 
the  'valve  openings  whicli  are  uncovered.  It  is  seen  that  the 
air  leaves  first  in  this  case;  the  water  is  struck  by  the  bucket 
surface  aftd  will  cause  considerable  shock  if  the  pump  is  running 
too,  rapidly. 

To  do  away  with  shock  and  to  decrease  valve  resistance,  the 
Edwards  Air  Pump,  Fig.  366,  was  introduced.  In  thisvair  pump 
water  and  air  enter  the  space  A  at  the  bottom  of  the  pump 
which  is  made  conical  in  form.  The  piston  B  which  is  driven 
from  the  steam  piston  in  /  by  two  rods  CC  extending  over 
the  shaft  and  crank,  is"  provided  with  a  conical  bottom.  As 
this  piston  descends  there  is  a  vacuum  produced:,  so  that  when 
the  top  of  the  piston  uncovers  the  openings  EE,  air  enters  from 
the  space  A  around  the  cylinder  barrel,  and  as  the  conical  bottom 
enters  the  water  in  the  bottom  of  A,  this  is  forced  around  the 
curved  passage  and  discharged  into  the  openings  at  E.  This 
continues  even  after  the  piston  starts  up,  as  the  momentum  of 
the  water  continues  its  motion.  This  discharge  of  water  into 
the  openings  as  the  piston  is  moving  upward  acts  as  a  valve  to 
keep  the  air  from  coming  ouTas  the  piston  ascends.  In  a  short 


462 


PUMPING   MACHINERY 


time,  however,  the  piston  covers  the  ports  or  openings  E  and 
then  the  air  and  water  are  compressed  until  the  pressure  is 
sufficient  to  overcome  the  atmospheric  pressure  on  the  head 
valves  at  H,  which  are  drowned  by  the  use  of  a  lip  around  the 
valve  deck.  The  piston  rods  CC  are  carried  through  long-sleeve 


FIG.  366. — Edwards  Air  Pump. 

s tuning  boxes  so  arranged  that  the  point  H,  at  which  leakage 
could  occur,  is  water  sealed,  leaving  only  one  stuffing  box  at 
the  plate  K  to  care  for.  This  is  a  simple  matter. 

The  Mullen  Valveless  Air  Pump,  Fig.  367,  is  somewhat 
similar  to  the  Edwards  Pump.  In  this  case,  the  deep  piston 
C,  provided  with  a  number  of  packing  grooves,  uncovers  ports 


SPECIAL  PUMPING  MACHINERY 


463 


4G4  PUMPING  MACHINERY 

at  the  center  of  the  cylinder.  The  vacuum  formed  by  the  motion 
of  the  piston  from  the  end  draws  in  the  air  and  water  from 
the  space  A  around  the  cylinder  barrel  and  the  return  of  the 
piston  cuts  these  ports  off  and  compresses  the  air  against  the 
spring  valves  at  the  end,  which  open  after  the  atmospheric 
pressure  is  reached.  The  spaces  BB  unite  and  lead  to  the  hot 
well. 

The  piston  of  this  pump  is  of  such  a  length  that  when  the 
ports  of  A  are  completely  uncovered,  the  piston  just  reaches 
the  end  of  its  travel,  leaving  a  small  amount  of  clearance  which 
is  filled  with  water,  so  that  as  soon  as  the  piston  leaves  the  end 
of  its  stroke  the  pressure  falls.  .  This  action  is  the  same  in  all 
wet  air  pumps.  The  piston  rod  of  the  Mullen  Air  Pump  is 
attached  to  a  cross  head  E  which  is  connected  to  the  cross 
head  -F  of  the  steam  end  by  two  rods,  so  that  the  shaft  and  crank 
may  be  cleared.  The  stuffing  box  at  D  is  water  sealed  to  cut 
down  the  leakage  of  air.  Air  leaks,  even  the  smallest,  are  to  be 
avoided  on  account  of  the  low  pressure  in  the  pump. 

Dry  air  pumps  have  become  quite  common.  They  were 
first  introduced  for  barometric  condensers  and  afterwards  for 
use  with  surface  condensers.  As  a  type  of  this  class,  the  Alberger 
Air  Pump,  Fig.  368,  is  shown.  In  it  the  piston  A  travels  from 
end  to  end  and  is  so  finished  on  the  ends  that  its  clearance  is 
small.  A  rotary  valve  below  the  spring  discharge  valve  D 
directs  the  air  to  the  suction  chamber  C  or  to  the  discharge. 
This  valve  B  is  positively  driven.  As  shown  in  the  left  hand 
figure,  the  air  on  the  right  is  being  drawn  from  C.  The  air  on 
the  left  is  not  discharged  until  it  has  reached  a  pressure  slightly 
above  the  atmosphere,  when  it  can  raise  the  spring  valve  D  and 
escape.  From  this  point  the  air  is  driven  out  as  the  piston 
continues  to  move  to  the  left.  When  the  end  of  the  stroke  is 
reached  the  pressure  on  one  end  is  atmospheric  and  that  on 
the  other  is  that  of  the  maximum  vacuum  carried  on  the  con- 
denser. The  piston  will  draw  no  more  on  this  stroke  and  the 
vacuum  of  the  condenser  would  not  be  effected  if  air  was  allowed 
to  enter  this  side  of  the  piston.  The  other  side  of  the  piston 
has  air  at  atmospheric  pressure,  filling  the  clearance  volume, 


SPECIAL  PUMPING  MACHINERY 


465 


including  the  passage  E.  If  this  air  remained  in  the  clearance 
volume,  it  would  cut  down  the  amount  of  air  drawn  in,  as  none 
would  enter  until  it  had  reached  the  pressure  of  the  condenser, 
if  the  suction  had  a  valve;  or,  it  would  change  the  vacuum  if 
allowed  to  discharge  back  into  the  condenser.  To  obviate  this, 
the  Alberger  Company  introduced  a  small  cross  connection 
passage  in  their  rotating  valve  which  connects  the  two  ends  of 
the  cylinder  when  the  piston  is  just  at  its  dead  point.  This 
arrangement  is  shown  on  the  right  of  Fig.  368,  where  the  piston 


FIG.  368. — Alberger  Air  Pump. 

is  at  the  right  hand  end  of  its  stroke.  In  this  position  the  space 
to  the  right  of  the  piston,  as  far  as  the  valve  B,  is  filled  with 
air  at  atmospheric  pressure,  while  the  space  on  the  left  of 
the  piston  is  at  condenser  pressure  and  has  reached  the  point 
where  it  will  not  receive  any  more  air  from  the  condenser. 
If  now  the  valve  cuts  off  both  of  these  ends  from  the  air  and 
condenser  but  connects  them  through  the  small  passage  in  the 
valve,  the  pressure  on  the  right  will  be  reduced  very  materially, 
as  the  volume  of  the  left  hand  end  of  the  cylinder  is  so  large 
compared  with  the  volume  of  the  clearance.  The  pressure  in 


466 


PUMPING   MACHINERY 


the  clearance  is  then  reduced  to  vacuum  pressure  when  the 
valve  opens  the  right  side  to  the  condenser,  and  the  left  hand 
end  of  low  vacuum  has  received  a  little  more  air  to  be 
discharged. 

This  valve  passage  therefore  makes  the  volumetric  efficiency 
of  the  pump  greater,  although  the  power  required  to  drive  the 

pump  for  a  given  quantity  is 
not  changed.    Fig.  369  shows 
the  card  from  such  a  pump 
compared  with  one  without 
this  arrangement.    The  solid 
.    curve  shows  how  the  clearance 
FIG.  369.— Alberger  Air  Pump  Card,     effect  is  reduced  and  how  the 

'    quantity    of    air    handled    is 

made  greater.  The  curve  also  shows  that  the  power  required 
in  compressors  of  the  same  displacement  is  increased.  The 
volume  of  air  handled  per  stroke  is  increased  from  AB  to  CB. 


FIG.  370. — Sewage  Pump. 

The  small  curve  at  C  is  due  to  the  pressure  not  falling  to  the 
vacuum  of  the  condenser  when  the  cross  connection  is  made, 
while  the  rise  in  pressure  at  B,  giving  a  starting  point  of  com- 


SPECIAL  PUMPING  MACHINERY  467 

pression  above  the  vacuum  pressure,  is  due  to  this  cross 
connection. 

Fig.  370  illustrates  a  pump  built  by  the  Laidlaw,  Dunn, 
Gordon  Co.  for  the  pumping  of  sewage.  The  valves  are  made 
large  on  account  of  the  solid  matter  in  suspension,  and  they  are 
placed  in  large  valve  boxes  on  the  sides  of  the  pump.  The 
pump  is  otherwise  similar  to  any  duplex  compound  pump. 

The  Underwriters'  Pump,  or  better,  the  "  National  Standard 


FIG.  371. — Underwriters'  Pump. 

Fire  Pump,"  to  which  attention  has  been  called  in  Chapter  II,  is 
illustrated  again  in  Fig.  371.  This  figure  illustrates  the  form 
used  in  accordance  with  the  specifications  of  the  National 
Board  of  Fire  Underwriters,  and  to  give  the  important  parts 
of  these  specifications  the  following  has  been  taken  from  their 
pamphlet. 


468  PUMPING  MACHINERY 

"  SPECIFICATIONS  OF  THE  NATIONAL  BOARD  OF 
FIRE  UNDERWRITERS  FOR  THE  MANUFACTURE 
OF  STEAM  FIRE  PUMPS,"  EDITION  OF  1908. 

UNIFORM  REQUIREMENTS. 

The  following  specifications  for  the  manufacture  of  Steam  Fire  Pumps, 
developed  from  those  originally  drawn  by  Mr.  John  R.  Freeman,  are  now 
used  throughout  the  whole  country,  having  been  agreed  upon  in  joint  confer- 
ence by  representatives  of  the  different  organizations  interested  in  this  class 
of  work.  They  will  be  known  as  "The  National  Standard,"  and  have  been 
up  to  this  time  adopted  by  the  following  associations: 

Associated  Factory  Mutual  Fire  Insurance  Companies. 

National  Board  of  Fire  Underwriters. 

National  Fire, Protection  Association,,'   ' 

NATIONAL    STANDARD    SPECIFICATIONS    FOR    THE    MANU- 
TURE   OF   STEAM    FIRE    PUMPS 

1.  Workmanship,      a.  The  general    character  and  accuracy  of  foundry 
and  machine  work  must  "throughout '  equal  that  of  the  best  steam-engine 
practice  of  the  times,  as  illustrated  in  commercial  engines  of  similar  horse- 
power. 

2.  Duplex  Only.     a.  I  Only  "  Standard  Duplex  pumps"  are  acceptable. 

So-called  "Duplex"  pumps,  consisting  of  a  pair  of  pumps  with 
"steam-thrown  valves"  actuated  by  supplemental  pistons,  are  not 
acceptable. 

Further,  the  direct-acting  duplex  has  the  great  advantage  over  a 
fly-wheel  pump  of  not  suffering  breakage  if  water  gets  into  steam  cylinder. 

3.  Sizes  of  Pumps,     a.  Only  the  four  different  sizes  given  in  the  table  on 
page  469  will  be  recognized  for  "National  Standard"  pumps. 

b.  The  tabular  sizes  of  steam  and  water  cylinders  and  length  of  stroke 
have  given  general  satisfaction  and  will  now  be  considered  as  standard. 

A  steam  piston  relatively  larger  than  necessary  is  a  source  of  weak- 
ness. It  takes  more  volume  of  steam,  and  gives  more  power  with  which 
to  burst  something  if  the  throttle  is  opened  wide  suddenly  during 
excitement. 

It  has  been  common  to  make  all  fire  pumps  with  water  plunger  of  only 
one -fourth  the  area  of  steam  piston,  with  the  idea  that  the  pump  could 
thereby  be  more  readily  run  at  night,  when  steam  was  low.  The  capacity 
in  gallons  is  thus  reduced  25  per  cent  as  compared  with  a  3  to  i  plunger 
en  the  same  steam  cylinders. 

Often,  especially  with  large  pumps,  "4  to  i"  construction  is  a  mis- 
take, and -gives  no  additional  security,  although  the  pump  might  start 
and  give  a  few  puffs  with  30  Ibs.  of  steam  on  banked  fires;  because, 
if  any  pump  of  whatever  cylinder  ratio  draws  50  or  100  horse-power  of 
steam  from  boilers  with  dead  fires,  it  can  run  effectively  only  a  very 
short  time  (ordinarily,  perhaps,  3  to  5  minutes),  unless  fires  are  first 
aroused  to  make  fresh  steam  to  replace  that  withdrawn. 

Steam   pressures   stated   above   must   be   maintained   at   the  pump, 


SPECIAL  PUMPING  MACHINERY 
NATIONAL  STANDARD  PUMP  SIZES 


469 


Pump  Sizes. 

1 

Capacity  at  100  Ibs., 
at  Pump. 

Boiler  Power 
Required. 

Full  Speed. 

c 
o 
tfi 

d 

1 

ft 

d 

X 

a 

tl 

1 

o 

CD  4 

g 

o3 

**"!  . 

0 

M 

•    J2 

a^j 

u- 

M    75 

a 

13 

~ct 

"o  w' 

t&    . 

& 

t-.  a 

,—  .   ^ 

M  i-, 

o 

PL!  d 

Id 

£-2 

Steam. 

Water. 

Stroke. 

About 

jl 

Jl 

I3 

1 

|J 

02 

Revolu 
minu 

Piston  ' 
minu 

14    X    7    XI2 

4  to  i 

Two 

500 

483 

IOO 

40 

70 

140 

14  x  7^x12 

520 

i  6  xg     xi  2 

3  to  i 

Three 

75° 

806 

US 

45  ' 

70 

140 

l8     X2O    XI2 

3  to  i 

Four 

loop 

999 

15° 

45 

70 

140 

i8ixiojxi2 

1050 

20    XI2     Xl6 

2j  tO    I 

Six 

1500 

1655 

200 

5° 

60 

160 

to  give  full  speed  and  100  Ibs.  water  pressure.  Pressure  at  boilers 
must  be  a  little  more,  to  allow  for  loss  of  steam  pressure  between  boiler 
and  pump.  Pumps  in  poor  order,  or  too  tightly  packed,  will  require 
more  steam. 

c.  Two  hundred  and  fifty  gallons  per  minute  is  the  standard  allowance 
for  a  good  i§  inch  (smooth  nozzle)  fire  stream. 

A  so-called  "Ring  Nozzle"  discharges  only  three-fourths  as  much 
water  as  a  smooth  nozzle  of  the  same  bore,  and  is  not  recommended. 

From  fifteen  to  twenty  automatic  sprinklers  may  be  reckoned  as 
discharging  about  the  same  quantity  as  a  i^-inch  hose  stream  under  the 
ordinary  practical  conditions  as  to  pipes  supplying  sprinkler  and  hose 
systems  respectively. 

4.  Capacity,  a.  Plunger  diameter  alone  will  not  tell  how  many  gallons 
per  minute  a  pump  can  deliver,  and  it  is  not  reasonable  to  continue  the  old 
time  notion  of  estimating  capacity  on  the  basis  of  100  feet  per  minute  piston 
travel. 

b.  The  capacity  of  a  pump  depends  on  the  speed  at  which  it  can  be  run, 
and  the  speed  depends  largely  on  the  arrangement  of  valves  and  passageways 
for  water  and  steam. 

c.  It  is  all  right  to  run  fire-pumps  at  the  highest  speed  that  is  possible 
without  causing  violent  jar,  or  hammering  within  the  cylinders.  Considerations 
of  wear  do  not  affect  the  brief  periods  of  fire  service  or  test,  hence  these  speeds 
are  greater  than  allowable  for  constant  daily  duty. 

d.  Careful  experiments  on  a  large  number  of  pumps  of  various  makes 
at  full  speed,  show  that  in  a  new  pump  with  clean  valves,  and  an  air-tight 
suction  pipe,  and  less  than  15  feet  lift,  the  actual  delivery  is  only  from  i| 
to  5  per  cent  less  than  plunger  displacement.     This  slip  will  increase  with 


470  PUMPING  MACHINERY 

wear,  and  for  a  good  average  pump  in  practical  use,  probably  10  per  cent  is 
a  fair  allowance  to  cover  slip,  valve  leakage,  slight  short-stroke,  etc. 

e.  Largely  from  tests,  but  partly  from  " average  judgment,"  and  recogniz- 
ing that  a  long  stroke  pump  can  run  at  a  higher  rate  of  piston  travel  in  lineal 
feet  per  minute  than  a  short  stroke  pump,  and  that  a  small  pump  can  make 
more  strokes  per  minute  than  a  very  large  one,  the  speeds  given  in  the  preceding 
table  have  been  adopted  as  standards  in  fire  service  for  direct-acting  (non- 
fly-wheel)  steam  pumps,  which  have  the  large  steam  and  water  passages 
herein  specified. 

/.  Rated  capacity  is  to  be  based  on  the  speed  in  the  preceding  table,  correct- 
ing the  plunger  displacement  for  one-half  the  rod  area  and  deducting  10  per 
cent  for  slip,  short -stroke,  etc. 

5.  Capacity  Plate,  a.  Every  steam  fire-pump  must  bear  a  conspicuous 
statement  of  its  capacity  securely  attached  to  the  inboard  side  of  air  chamber, 
thus: 

NATIONAL  STANDARD  FIRE  PUMP 
16  X  9  X  12 


CAPACITY 

750  GALLONS  PER  MINUTE,  OR  THREE  GOOD  l£-IN.  SMOOTH  NOZZLE 

FIRE  STREAMS 


FULL    SPEED 

70    REVOLUTIONS    PER    MINUTE 


Never  let  steam  get  below  50  pounds,  nights,  Sundays,  or  at 
any  other  time 

b.  This  plate  must  have  an  area  of  not  less  than  one  square  foot,  and 
must  be  made  of  an  alloy  at  least  85  per  cent  aluminum  and  the  remainder 
zinc.    The  letters  must  be  at  least  one-half  inch  in  height,  plain  and  distinct, 
with  their  surfaces  raised  on  a  black  background  and  buffed  off  to  a  dead 
smooth  finish. 

c.  A  smaller  plate  of  composition  must  be  attached  to  steam  chest  bearing 
the  size  of  pump,  the  shop  number,  and  the  name  of  shop  in  which  the  pump 
was  built. 

6.  Strength  of  Parts,     a.  The  maker  must  warrant  each  pump  built  under 
these  specifications  to  be,  at  time  of  delivery,  in  all  its  parts  strong  enough 
to  admit  of  closing  all  valves  on  water  outlet  pipes  while  steam  valve  is  wide 
open  and  steam  pressure  eighty  pounds,  and  agree  to  so  test  it  before  shipment 
from  his  works. 

b.  The  pump  must  be  warranted  so  designed  and  with  such  arrangement 
of  thickness  of  metal  that  it  shall  be  safe  to  instantly  turn  a  full  head  of  steam 
on  to  a  cold  pump  without  cracking  or  breaking  the  same  by  unequal  expansion. 

7.  Shop   Inspection.    A  systematic  shop  inspection   must   be  given   to 
each  pump  to  ensure  complete  workmanship,  and  to  prevent  the  use  of  defective 
parts,  improper  materials,  or  the  careless  leaving  of  foreign  matter  in  any 
part  of  the  cylinders  or  chests. 


SPECIAL  PUMPING  MACHINERY  471 


THE  STEAM  END. 

8.  'Steam  Cylinders,  a.  These  must  be  of  hard  close  iron  with  metal 
so  distributed  as  to  insure  sound  castings  and  freedom  from  shrink  cracks. 
The  following  are  the  minimum  thicknesses  acceptable: 


14"  Diam.       I"  thick. 


1  6 


" 


1 8"  Diam.      i"  thick. 


20 


b.  The  inside  face  of  the  steam  cylinder  heads  and  the  two  faces  of  the 
piston  must  be  smooth  surfaces,  fair  and  true,  so  that  if  the  piston  should  hit 
the  heads  it  will  strike  uniformly  all  around,  thus  reducing  to  a  minimum 
the  chances  of  cramping  the  piston  rod  or  injuring  the  pump. 

c.  All  flanged  joints  for  steam  must  be  fair  and  true  and  must  be  steam 
tight  under  80  pounds  pressure  if  only  a  packing  of  oiled  paper  i-ioo  inch 
thick  covered  with  graphite  were  used.     Jenkins,  " Rainbow"  or  equivalent 
packing  of  not  exceeding  ^  inch  original  thickness  is  acceptable.      Oiled 
paper  is  not  acceptable  as  a  final  packing,  as  it  burns  out. 

For  size  of  steam  and  exhaust  pipes,  standard  flanges  and  bolting,  see 
Art.  39. 

d.  Heads  at  both  ends  of  cylinder  must  be  beveled  off  very  slightly  over 
a  ring  about  one  inch  wide,  or  equivalent  means  provided  to  give  steam  a 
quick  push  at  piston,  should  it  stand  at  contact  stroke. 

The  specifications  originally  required  machine  facing  for  all  these 
surfaces.  The  art  of  machine  molding  from  metal  patterns  with  draw 
plates,  etc.,  has,  however,  attained  such  excellence  in  certain  shops, 
that  in  regular  practice  "foundry  faced"  cylinder  heads  and  piston 
faces  can  be  made  true  and  fair,  and  steam  joints  can  be  made  tight 
under  80  Ibs.  pressure  with  a  packing  of  oiled  paper  only,  i-ioo-inch 
thick. 

11.  Steam  Ports,     a.  The  area  of  each  exhaust  steam  passage,   at  its 
smallest  section,  must  not  be  less  than  4  per  cent  of  the  area  of  the  piston 
from  which  it  leads. 

b.  Each  admission  port  must  be  not  less  than  2^  per  cent  of  area  of  its 
piston  and  to  avoid  wasteful  excess  of  clearance,  these  passages  should  not 
be  bored  out  larger  in  interior  of  casting  than  at  ends  or  passage. 

c.  The  edges  of  the  steam-valve  ports  must  be  accurately  milled,  or  chipped 
and  exactly  filed  to  templets,  true  to  line,  and  the  valve  seat  must  be  accurately 
fitted  to  a  plane  surface,  all  in  a  most  thorough  and  workmanlike  manner  and 
equal  to  high-grade  steam-engine  work. 

d.  To  guard  against  a  piston  ring  catching  in  the   large  exhaust  ports, 
these  ports  must  have  a  center  rib  with  cylinder  at  cylinder  wall.     See  also 
Art.  13  d. 

12.  Steam-clearance  Space,     a.  Clearance  (including  nut-recess,  counter- 
bore,  and  valve  passages)  must  not  exceed  5  per  cent  for  contact  stroke  or 
about  8  per  cent  for  nominal  stroke  (i.e.,  contact  stroke  should  overrun  nominal 
stroke  at  each  end  about  one-half  inch). 

b.  The  clearance  space  between  face  of  piston  and  cylinder  head  must 
be  reduced  to  smallest  possible  amount,  and  these  contacting  surfaces  be 


472  PUMPING  MACHINERY 

flat,  without  projections  or  recesses  other  than  the  piston  rod  nut  and  its 
recess. 

Some  makers,  with  the  idea  that  a  fire  pump  need  not  be  economical, 
have  not  taken  pains  to  keep  these  waste  spaces  small. 

Securing  small  clearance  costs  almost  nothing  but  care  in  design,  and 
is  often  of  value,  since  at  many  factories  boiler  capacity  is  scant  for  the 
large  quantity  of  steam  taken  by  a  fire  pump  of  proper  size. 

13.  Steam  Pistons,     a.  May  be  either  built  up  or  colid,  as  maker  thinks 
best. 

It  is  believed  that  "solid"  (cored)  pistons  with  rings  "sprung  in," 
are  for  fire  pumps  much  preferable  to  built-up  pistons,  since  follower 
bolts  do  sometimes  get  loose. 

b.  The  thickness  of  piston  should  be  about  one-fourth  of  its  diameter. 
If  solid,  walls  should  be  not  less  than  \  inch  thick,  and  special  care  should 
be  given  to  shop  inspection  to  insure  uniformity  of  thickness. 

This  will  demand,  for  the  four  sizes  of  pumps,  pistons  as  follows: 
5oo-gal.  75o-gal.  xooo-gal.  isoo-gal. 

Diameter  14  in.     Diameter  16  in.          Diameter  18  in.      Diameter  20  in. 
Thickness  33  in.     Thickness    4  in.          Thickness  4^  in.     Thickness     5  in. 

Manufacturers  desiring  to  use  existing  patterns  approximating 
these  thicknesses  may  be  allowed  to  do  so  after  due  consideration  of 
working  drawings. 

c.  If  built-up  pistons  are  used,  involving  follower  bolt3,  such  bolts  must 
be  of  best  machinery  steel,  with  screw  thread  cut  for  nbout  twice  the  diameter 
of  the  bolt  and  fitting  tightly  its  whole  length. 

d.  The  width  of  each  piston  ring  must  exceed  the  length  of  the  large 
exhaust  port  by  at  least  \  inch. 

This  is  to  avoid  the  possibility  of  piston  ring  catching  in  the  port. 
See  also  Art.  n  d. 

14.  Steam  Slide-valves,      a.  Slide  valves  must  be  machine  fitted  on  all 
four  of  the  outer  edges,  the  exhaust  port  edges,  and  the  surfaces  in  contact 
with  rod  connections. 

b.  The  slide  valve  itself  must  have  its  steam  and  exhaust  edges  fitted  up 
"  line  and  line"  with  their  respective  steam  and  exhaust  ports. 

The  adding  of  lap  to  these  edges  in  lieu  of  lost  motion  is  not  accept- 
able further  than  a  possible  •$$  of  an  inch  to  cover  inaccuracies  of  edges. 

c.  The  valves  must  be  guided  laterally  by  guide  strips  cast  in  steam  chest, 
and  these  strips  must  be  machine  fitted.     The  lateral  play  at  these  surfaces 
should  not  exceed  &  inch.      The   height  of   these  guide  strips  should  not 
be  less  than  \  inch,  measuring  from  valve  seat. 

The  construction  must  be  such  as  to  absolutely  preclude  the  possibility 
of  the  valve  riding  up  on  top  of  this  guide  strip. 

d.  The  valves  must  be  guided  vertically  by  the  valve-rod  itself,  the  inside 
end  of  which  must  be  kept  in  alignment  by  the  usual  form  of  tail-rod  guide. 

The  vertical  play  at  these  parts  should  not  exceed  |  of  an  inch. 

e.  The  surface  of  valves  must  be  machine  faced  and  accurately  fitted  to 
a  plane  surface,  and  be  steam  tight  when  in  contact  with  the  seat  of  steam  valve. 

15.  Steam  Slide-valve  Adjustment,     a.  The  lost  motion  at  the  valves  and 


SPECIAL  PUMPING  MACNINERY  4>3 

the  settling  of  them  must  be  determined  by  a  solid  hub  on  the  rod,  finished 
in  the  pump  shop  to  standard  dimensions,  so  that  no  adjustment  is  possible 
after  the  pump  is  once  set  up. 

It  is  recognized  that  the  practice  of  making  adjustable  valve  tappets 
located  outside  of  the  steam  chest  is  a  good  thing  in  a  large  pump  in 
constant  service  and  operated  by  a  skilled  engineer,  but  for  the  infre- 
quently used  ordinary  fire  pump,  the  utmost  simplicity  is  desirable, 
and  it  is  best  not  to  tempt  the  ordinary  man  to  readjust  the  valve  gear. 

The  common  form  of  lost  motion  adjustment  consisting  of  nut  and 
check  nut  at  each  end  of  the  slide  valve  is  not  acceptable,  as  these  nuts 
are  liable  to  become  loose  and  may  be  incorrectly  reset  by  incompetent 
persons.  A  long  rectangular  nut  in  the  center  of  the  valve  is  also  not 
acceptable,  as  it  can  be  moved  out  of  adjustment.  A  solid  hub  made  as 
a  part  of  the  rod  is  required,  as  it  absolutely  avoids  the  possibility  of 
the  hub  becoming  loose,  an  accident  possible  with  a  separate  hub  attached 
to  the  rod. 

The  amount  of  lost  motion  should  generally  be  such  that  admission 
takes  place  at  about  f.  of  the  stroke  of  the  piston,  i.e.,  for  1 2-inch  stroke 
R.H.  valve  will  be  about  to  open  when  L.H.  piston  has  moved  7$  inches 
to  8  inches  from  the  beginning  of  stroke.  When  piston  is  at  end  of 
stroke  the  ports  should  bs  full  open. 

1 6.  Rock  Shafts,    Cranks,   Links,    etc.      a.  Rock  shafts  must  be  either 
forged  iron,  forged  steel,  cr  cold  rolled  steel.     Cast  iron  is  not  acceptable: 
The  following  are  th^  minimum  diameters  acceptable: 

500  gallon  pump i^  in. 

750  gallon  pump if  in. 

1000  gallon  pump 2  in. 

1500  gallon  pump 2  to  2  J  in. 

b.  The  rock-shaft  bearings  must  be  bushed  with  bronze  and  the  bushings 
pinned  firmly  in  place.     The  length  of  each  of  these  non-corrosive  bearings 
must  be  not  less  than  4  inches. 

c.  Rock-shaft  cranks,  valve-rod  heads,  valve-rod  links,    and   piston-rod 
spools  or  cross  heads  may  be  wrought  iron  or  steel  forgings,  or  steel  castings. 
If  of  a  heavy,  strong  pattern,   these  parts,  with  the  exception  of  valve-rod 
links,  may  be  of  semi-steel  or  cast  iron. 

d.  The  sectional  area  of  all  connections  between  rock-shaft  cranks  and 
valve  rod  must  be  such  as  to  give  a  tensile  or  compressive  strength  substantially 
equal  to  that  of  the  valve  rod. 

'  17.  Valve-motion  Levers,  a.  The  valve-motion  levers  must  be  steel, 
wrought  iron,  or  steel  castings.  Cast  iron  is  not  acceptable.  Steel  castings, 
if  used,  must  be  deeply  stamped  with  the  name  of  the  makers,  with  letters 
one-eighth  inch  high,  near  the  upper  end  of  each  lever,  where  it  can  easily 
be  seen, — thus  " — Steel  Castings." 

Cast-iron  arms,  if  bulky  enough  to  be  safe  against  external  blows, 
are  awkward  in  shape.  The  sectional  area  necessary  for  any  arm 
depends  upon  the  means  provided  for  preventing  a  sidewise  stress  on 
the  lever,  due  to  rotation  of  piston  or  friction  of  its  connection  to  piston 
rod.  The  spool  or  cross  head  on  the  piston  rod  should  be  so  designed 
that  no  sidewise  strain  can  be  thus  produced  in  the  lever. 

b.  The  levers  must  have  a  double  or  bifurcated  end  at  cross  head. 

The  double  end  is  less  likely  than  a  single  end  to  put  undue  stress 


474 


PUMPING  MACHINERY 


on  the  lever  as  the  rod  turns,  and  is  also  less  likely  to  give  trouble  from 
lack  of  lubrication  or  from  a  loosening  of  any  small  parts,  and  has 
proved  to  be  the  most  satisfactory  arrangement. 

18.  Valve-motion  Stand.      a.  The  valve  motion  stand  must  be  securely 
dowel -pinned  to  the  yoke  castings,  to  prevent  any  movement  after  once  adjusted. 

19.  Cushion    Valves,     a.  Cushion -release  valves  regulating  the  amount 
of  cushion  steam  retained  at  ends  of  stroke  must  be  provided. 

b.  The  cushion  release  must  be  through  an  independent  port,  as  shown 


FIG.  372.  —  Cushion  Valve. 

in  Fig.  372,  so  located  as  to  positively  retain  a  certain  amount  of  cushion  steam. 

The  old  form  of  cushion  release  through  bridge  between  ports  is 
not  acceptable.  This  form,  while  leading  into  the  exhaust  passage  as 
formerly,  differs  by  starting  from  a  small  independent  port  (about 
^-inch  wide  by  z\  inches  long)  through  the  cylinder  wall,  located  about 
|  or  \  inch  back  from  the  cylinder  head.  (The  exact  position  for  affording 
the  best  action  has  to  be  determined  by  experiment  with  each  different 
make  of  pump,  as  it  depends  somewhat  on  the  extent  of  clearance  space 
and  on  the  point  of  closure  of  exhaust  by  piston  and  somewhat  on  the 
weight  of  reciprocating  parts)  . 

This  style  of  cushion  port  makes  the  pump  safer  in  case  cushion 
valves  are  unskilfully  left  open  too  wide,  and  tends  to  prevent  a  pump 
from  pounding  itself  to  pieces  in  case  of  a  sudden  release  of  load,  as  by 
a  break  in  suction  or  delivery  mains,  or  by  a  temporary  admission  of 
air  to  suction  pipe. 

Pumps  made  with  this  form  of  cushion  release  have  given  very 
satisfactory  results,  and  if  the  ports  are  properly 
located,  there  will  be  no  rebound  of  piston. 

c.  Cushion  valves    must    be  .  always    provided 
with  hand-wheels  marked  as  per  sketch,  Fig.  373, 
for    the  reason    that    a   very  few 
men  in  charge  of  fire  -pumps  are 
found  to  clearly  understand  or  to 
remember  their  use. 

The    lettering    must    be   very 
open,  clear  and  distinct,  not  liable 
to  be  obscured  by  grease  and  dirt, 
and  of  a  permanent  character. 

It  is  desirable  that  spindle  or  wheel  be  so  formed  that 
a  monkey  wrench  can  get  a  grip  to  open  a  jammed 
valve.  Fig.  374  shows  the  stem  flattened  for  this  purpose.  pIGt 

d.  The  valve  and  stem  of  cushion  valve  must  be  in 
one  piece  without  any  swivel  joint. 

Swivel  joints  are  apt  to  come  apart  and  make  it  impossible  to  operate 
the  valve. 


FIG.  373.— Hand  Wheel. 


_  Flat- 
tened  Stem. 


3PECIAL  PUMPING  MACHINERY 


475 


20.  Piston  Rods.  a.  Piston  rods  for  their  entire  length  must  be  of  solid 
Tobin  Bronze,  and  the  distinguishing  brand  of  the  manufacturers  of  this 
metal  must  be  visible  on  at  least  one  end  of  each  rod. 

b.  The  sizes  must  be  not  less  than  in  table  below. 


Size  of  Pump. 

500  gal. 

75°  gal. 

1000  gal. 

1500  gal. 

Diameter  of  rod  

2  inch 

2\  inch 

2f  inch 

2|  inch 

c.  The  size  and  form  of  connection  of  rod  to  "piston    plunger  and  cross 
head  must  be  such  that  the  stress  in  pounds  per  square  inch  at  bottom  of 
screw  thread,  or  at  such  other  point  of  reduced  area  as  receives  the  highest 
tensile  stress,  shall  not  exceed  8000  pounds  per  square  inch,  when  the  steam 
pressure  acting  on  the  piston  is  80  pounds  per  square  inch. 

d.  Piston  rod  nuts,  in  both  steam  and  water  ends,  must  be  tightly  fitted, 
and  preferably  of  a  finer  thread  than  the  United  States  Standard.    This  is  to 


FIG.  375. — Lock  Nut. 

avoid  as  much  as  possible  the  unnecessary  weakening  of  the  rod  at  the  bottom 
of  the  thread,  and  to  reduce  the  tendency  of  the  nut  to  work  loose. 

In  practice  eight  threads  per  inch  has  been  found  to  give  good 
satisfaction. 

e.  In  addition  to  a  tightly  fitting  nut,  some  reliable  device  must  be  provided, 
in  both  steam  and  water  ends,  for  absolutely  preventing  these  nuts  from 
working  off. 

Fig.  375  shows  one  form  of  such  a  locking  device  and  illustrates 
the  kind  of  security  desired. 

This  device  combines  the  advantage  of  a  taper  key  and  a  split  pin, 
and  the  elongated  key-slot  gives  sufficient  leeway  to  always  insure  that 
the  key  can  be  driven  up  tight  against  the  nut  and  thus  prevent  it  from 
even  starting  to  work  off.  Other  methods  will  be  approved  in  writing, 
if  found  satisfactory. 

21.  Valve  Rods.  a.  Valve  rods  for  their  entire  length  must  be  of  solid 
Tobin  Bronze,  with  sizes  not  less  than  in  table  below. 


Size  of  Pump. 

500  gal. 

750  gal. 

1000  gal. 

1500  gal. 

Diameter  of  rod                     

i  inch 

i£  inch 

i£  inch 

i  -J-  inch 

b.  The  net  area  of  valve-rod  at  its  smallest  section  subject  to  tensile  stress, 


476  PUMPING  MACHINERY 

must  not  be  smaller  than  at  bottom  of  U.  S.  Standard  screw  thread  on  rod 
of  diameter  given  above. 

The  construction  of  this  rod  as  affecting  lost  motion  at  slide  valve 
is  specified  under  Art.  15. 

22.  Stuffing  Boxes,     a.  Ail  six  stuffing  boxes  must  be  bushed  at  the  bottom 
with  a  brass  ring  with  suitable  neck  and  flange,  and  the  follower  or  gland 
must  tc  either  of  solid  brass,  or  be  lined  with  a  brass 
shell   ^-inch  thick,  having  a  flange  next  the  packing, 
as  shown  in  the  sketch. 

The  bottom  of  stuffing  boxes  and  the  end  of  the 
glands  should  taper  slightly  towards  the  center,  as  per 
sketch. 

b.  These  glands  should  be  strong  enough  to  withstand 
considerable  abuse,  so  as  not  to  break  from  the  unfair 
BottU  treatment  of  unskilled  men. 

23.  Pressure  Gauge,     a.  A   pressure  gauge  of   the 

Lane  double  tube  spring  pattern  with  5-inch  case  must  be  provided  and 
attached  to  the  steam  chest  inside  the  throttle  valve. 

The  dial  of  gauge  should  be  scaled  to  indicate  pressures  up  to  120  pounds 
and  be  marked  "STEAM." 

24.  Drain  Cocks,    a.  Four  brass  drain  cocks,  each  with  lever  handle 
and  of  one -half  inch  bore,  are  to  be  provided,  and  located  one  on  each  end 
of  each  steam  cylinder. 

25.  Oiling  Devices,    a.  A  one-pint  hand  oil  pump,  to  be  connected  below 
the  throttle,  and  a  one-pint  sight  feed  lubricator,  to  be  connected  above  the 
throttle,  must  be  furnished  with  each  pump. 

b.  Oiling  holes  must  be  provided  for  all  valve  motion  pins,  and  for  each 
end  of  both  rock  shafts. 

26.  Stroke   Gauge,     a.' A   length -of -stroke  index    must   be   provided   for 
each  side  of  pump.    These  must  be  of  simple  form  for  at  all  times  rendering 
obvious  the  exact  length  of  stroke  which  each  piston  is  making,  and  thus 
calling  attention  to  improper  adjustments  of  cushion  valves  or  stuffing  boxes. 

b.  The  gauge  piece  over  which  the  index  slides  must  have  deep,  conspicuous 
marks  at  ends  of  nominal  stroke,  and  also  light  marks  at  extreme  positions; 
it  need  contain  no  other  graduations. 

c.  This  stroke  index  must  be  rigidly  secured  to  cross  head  in  such  a  way 
that  it  cannot  get  loose  or  out  of  adjustment. 

THE  WATER  END. 

27.  Water  Cylinders,     a.  These  must  be  of  hard  close  iron  with  metal 
so  distributed  as  to  ensure  sound  castings,  and  freedom  from  shrink  cracks. 

b.  The  design  should  be  along  lines  best  calculated  to  resist  internal 
pressures  so  as  to  avoid  as  much  as  possible  the  need  of  ribs  for  stiffening. 

c.  They  must  be  capable  of  withstanding,  without  showing  signs  of  weak 
ness,  the  pressures  and  shocks  due  to  running  under  the  conditions  mentioncu 
in  chapter  "Tests  for  Acceptance,"  Arts.  48-54. 


SPECIAL   PUMPING   MACHINERY 


477 


The  suction  chamber  should  be  able  to  withstand  a  water  pressure  of  100 
pounds. 

Although  suction  chambers  are  not  regularly  subject  to  a  pressure, 
it  is  sometimes  desired  to  connect  them  to  public  water  supplies,  and 
where  foot  valves  are  used  there  is  a  chance  of  getting  pressure  on  the 
suction,  so  that  ample  strength  is  necessary. 

Foundry  finish  may  be  permitted  on  the  joints  at  water  cylinder 
heads  and  at  hand-hole  plates,  provided  surfaces  are  so  true  that  a 
rubber  packing  not  over  ^  of  an  inch  in  thickness  is  sufficient  to  secure 
perfect  tightness. 

d.  Conveniently  placed  hand -holes  of  liberal  size  must  be  provided  for 
the  ready  examination  and  renewal  of  valve  parts  at  the  yoke,  end  of  water 
cylinders  and  in  the  delivery  chamber. 

This  will  necessitate  holes  not  less  than  6x8  inches,  or  its  equivalent, 
for  the  two  largest-size  pumps,  and  holes  proportionately  as  large  for 
the  500-  and  75o-gallon  pumps.  The  easy  access  to  the  valve  parts  is 
of  vital  importance,  and  must  receive  careful  attention. 

e.  The  thickness  of  metal  for  cylinder  shell,  valve  decks,  partitions,  ribs, 
etc.,  will  depend  largely  upon  the  form  of  construction,  but,  in  a  general  way, 
to  establish  safe  minimums  for  the  average  water  cylinder,  of  nearly  cylindrical 
form,  whose  flat  surfaces  are  stiffly  ribbed,  we  submit  the  table  belo.w: 


Size  of  Pump. 

500  gal. 

750  gal. 

1000  gal. 

1500  gal. 

Thickness  of  cylinder  shell  when 
of  nearly  cylindrical  form  
Thickness   of   valve   decks   when 
well  ribbed 

Inches 
1 
ll 

Inches 
I 
ii 

Inches 
I* 

il 

Inches 

*i 

ii 

Thickness     transverse     partition, 
depending  on  ribbing  

i  i  to  i  $ 

il  to  ii 

i£  to  2 

l£   tO   2 

Thickness    of   longitudinal    parti- 
tion  depending  on  ribbing 

T  1  tO   I  i 

1  1  to  1  5 

I  i  to   2 

I  i  to  2 

Thickness  of  ribs    
Thickness  of  suction  chamber.  .  .  . 
Thickness  of  delivery  chamber 

1 

1 
1 

i 
I 

i 

I 
I 
'I 

T 
I 

'i 

Lighter  construction  than  herein  specified  will  not  be  regarded  as  satis- 
factory, and  any  construction  will  be  finally  passed  upon  on  examination  of 
drawings. 

/.  The  bolting  of  all  parts  of  the  water  end  is  to  be  of  such  strength  that 
the  maximum  stress  at  bottom  of  screw  thread  will  not  exceed  10,000  pounds 
per  square  inch  (disregarding  for  the  moment  the  initial  stress  due  setting 
up  nuts)  for  a  water  pressure  of  200  pounds  per  square  inch,  computed  on 
an  area  out  to  center  line  of  bolts. 

No  stud  or  tap  bolt  smaller  than  f  inch  should  be  used  to  assemble  parts 
subject  to  the  Ltress  of  water  pressure,  as  smaller  bolts  are  likely  to  be  twisted 
off.  This  does  ret  apply  to  standard  flanges  where  through  bolts  are  used. 

Although  these  pumps  are  not  expected  to  be  designed  for  a  regular 


478  PUMPING   MACHINERY 

working  water  pressure  of  240  or  320  Ibs.,  it  is  expected  that  bolts, 
shells,  rods,  etc.,  will  be  figured  to  stand  this  comparatively  quiet, 
temporary,  high  pressure,  exclusive  of  further  allowance  for  initial 
stress  due  setting  up  of  bolts,  with  a  factor  of  safety  of  at  least  four. 

This  high  test  pressure  is  analogous  to  the  custom  of  proving  all 
common  cast-iron  water  pipes  to  300  Ibs.  and  all  common  lap-welded 
steam  pipes  to  500  Ibs.  per  square  inch,  and  common  water- works  gate 
valves  to  400  Ibs.,  even  though  these  are  to  be  regularly  used  at  much 
less  pressure. 

We  are  assured  that  castings  no  heavier  than  at  present  used  by  the 
best  makers  will  stand  this  test,  if  properly  shaped  and  liberally  bolted. 

g.  For  requirements  for  stuffing  boxes,  see  Art.  22. 

28.  Water -Plungers  and  Bushings,  a.  The  "inside  plunger  and  bushing" 
is  preferred  for  all  situations  where  the  water  is  free  from  grit  or  mud. 

b.  Water-plungers  must  be  of  solid  brass  or  bronze,  and  the  bushing  in 
which  they  slide  must  also  be  of  brass  or  bronze.     The  composition  of  the 
plunger  and  its  bushing  should  be  of  very  hard,  though  dissimilar  alloys, 
to  ensure  good  wearing  qualities. 

For  material  and  size  of  piston  rods  and  lock  for  nuts,  see  Art.  20. 

With  poor  alignment  or  bad  workmanship  or  lack  of  skill  in  mixing 
the  alloys,  brass  plungers  are  liable  to  score  and  give  trouble;  but  with 
proper .  selection  of  alloys  and  true  cylinders  accurately  aligned,  they 
can  be  made  to  run  all  right  wherever  iron  ones  can.  It  is  quite  a  fine 
point  to  get  these  wearing  surfaces  just  right;  and  this  is  wherein  the 
experience,  skill,  and  shop  practice  of  one  maker  is  likely  to  be  much 
superior  to  that  of  another  working  under  the  same  specification. 

c.  The  length  of  machined  cylindrical  bearing  within  the  partition  must 
be  not  less  than  two  inches.     The  plunger  bushings  must  have  a  faced  seat 
transverse  to  its  axis  against  partition,    forming  a  water-tight  ground  joint 
not  less  than  one-half  inch  wide. 

Any  rubber  gasket  or  other  compressible  packing  for  making  this  joint 
water-tight  is  not  acceptable. 

d.  The  construction  of  bushing  and  hole  in  partition  must  be  such  that 
a  cylindrical  shell  for  use  with  a  packed  piston  can  be  interchangeably  inserted 
in  its  place  and  secured  by  the  same  bolts. 

This  can  readily  be  arranged,  and  enables  a  packed  piston  to  be 
inserted  in  place  of  a  plunger  subsequent  to  the  installation  of  the  pump 
with  a  minimum  of  expense,  should  this  become  desirable  from  change 
of  conditions  at  any  future  time. 

e.  Small  transverse  grooves  cut  within  the  sliding  surface  of  the  plunger 
bushing,  with  a  view  to  lessen  the  leakage,  are  not  acceptable. 

Although  a  slight  advantage  in  this  respect  for  clean  water,  they  are 
a  disadvantage  on  the  whole,  as  dirt  catches  in  them  in  the  ordinary 
situation  and  cuts  the  plungers. 

29  Standard  Dimensions  of  Plungers  and  Plunger  Bushings.  a.  To 
bring  all  these  expensive  parts  to  the  same  standard  of  weight  and  bearing 
surface,  the  following  dimensions  are  specified  as  the  least  that  will  be  accept- 
able. These  are  based  on  a  length  of  plunger  which  uncovers  the  bushings 
one  inch  at  end  of  nominal  stroke; 


SPECIAL  PUMPING  MACHINERY 


479 


SOLID  BRONZE  PLUNGERS  AND  BUSHINGS 


Size  of  Pump. 

500  gal. 

750  gal. 

1000    gal. 

1500    pril. 

Plunger  — 

Inches. 

Inches. 

Inches. 

Inches. 

Diameter  

7  or  i\ 

9 

10  or  loj 

12 

Length 

17 

1  7 

18 

2.4. 

Thickness  of  transverse   parti- 

tion 

| 

• 

•4. 

3 

Thickness  next  to  partition  .... 

I 

1 

I 

I 

Thickness  next  to  end  

ft 

1 

f 

1 

Number  of  ribs          .    . 

4 

4 

6 

6 

Thickness  of  ribs 

A 

A 

| 

a 

Bushing  — 

Length 

7 

7" 

8 

10 

Thickness  at  end                 .... 

A 

£ 

| 

1 

Thence    tapered    evenly    to    a 

thickness  next  to  bearing  of 

not  less  than  

i 

f 

f 

1 

Thickness  at  the  center  bearing 

not  less  than  

I 

.    I 

1 

H 

30.  Water  Pistons  and  Bushings.  a.  The  " water  piston  with  fibrous 
packing"  is  preferred  for  many  situations  in  the  West  or  South,  or  for  water 
containing  grit  or  mud,  like  that  of  the  Ohio  River;  and,  for  the  comparatively 
few  cases  where  pump  pressure  governors  are  used,  the  packed  piston  will 
give  better  service  and  longer  wear. 

b.  The  removable  bushing  or  cylinder  in  which  this  piston  works  must  be 
of  solid  bronze. 

c.  As  stated  in  Art.  28  d,  this  bushing  should  be  so  constructed  as  to  be 
readily  interchangeable  with  the  bushing  of  the  inside  plunger  type. 

d.  The  length  of  cylindrical  bushing  must  be  such  that  the  outer  edge  of 
packing  will  come  short  of  the  edge  of  bushing  at  contact  stroke  about  \  inch 
and  not  uncover. 

e.  The  thickness  of  the  cylindrical  bushings  must  be  not  less  than  that 
given  in  the  following  table: 

BUSHINGS  FOR  PACKED  WATER  PISTONS 


Size  of  Pump. 

500  gal. 

750  gal. 

1000   gal. 

1500   gal. 

Solid  Bronze  — 

Inches. 

Inches. 

Inches. 

Inches. 

Thickness  at  extreme  end  

A 

\ 

i 

ft- 

Tapered  evenly  from  end  to  a 

thickness  next  to  bearing  of 

not  less  than  

A 

f 

H 

f 

Thickness  at  center  bearing,  at 

least              

1 

£ 

f 

ri 

480  PUMPING  MACHINERY 

f.  In  other  respects,  the  specifications  for  plunger  bushings,  already  given 
in  Art.  28,  will  apply  to  the  above. 

g.  The  water  piston  used  in  the  shell  described  above  must  expose  not 
less  than  2  inches  in  width  of  fibrous  packing,  and  must  be  of  bronze,  with 
disc  and  follower  accurately  turned  to  a  sliding  fit,  so  that  the  leakage  past 
it  will  be  a  minimum,  even  when  no  fibrous  packing  is  in  place.    There  must 
be  at  least  2  inches  in  length  of  metallic  bearing  on  both  disc  and  follower. 

The  follower  must  be  accurately  centred,  and  fitted  to  hub  of  piston,  so 
that  alignment  will  not  be  disturbed  if  taken  apart. 

h.  The  water  piston  must  be  of  simple  and  strong  construction,  with 
follower  bolts  tightly  fitted,  and  with  fibrous  packing  so  cut  as  to  prevent 
by-passing. 

i.  All  materials  used  in  construction  of  piston,  except  packing,  must  be 
brass,  bronze,  or  other  non-corrosive  metal. 

j.  Bushing  studs  must  be  of  Tobin  bronze,  and  of  such  size  and  number, 
that  the  maximum  stress  at  the  bottom  of  the  screw  thread  shall  not  exceed 
10,000  pounds  per  square  inch,  in  the  event  of  plunger  becoming  fast  in  the 
bushing  with  80  pounds  of  steam  in  the  steam  cylinders. 

k.  For  each  bushing  stud  there  must  be  provided  a  composition  nut  and 
check  nut. 

/.  All  minor  parts  exposed  to  the  action  of  water  in  water  cylinder,  that 
are  not  herein  specified,  must  be  of  brass,  bronze,  or  other  non-corrosive 
material. 

31.  Pump  Valves,     a.  All  the  suction  and  discharge  valves  in  any  one 
pump  must  be  of  the  same  size  and  interchangeable. 

b.  There  must  be  a  clear  space  around  each  rubber  valve,  between  it 
and  the  nearest  valve,  equal  to  at  least  one-fourth  of  the  diameter  of  the  valve, 
or  between  it  and  the  wall  of  the  chamber  of  at  least  one-eighth  of  the  diameter 
of  the  valve. 

c.  These  valves  must  be  of  the  very  best  quality  of  rubber,  of  medium 
temper   with  a  face  as  soft  as  good  wearing  quality  will  permit. 

They  must  be  double-faced,  so  they  can  be  reversed  when  one  face  is 
worn. 

The  quality  of  rubber  is  almost  impossible  of  determination  by 
brief  inspection  or  by  chemical  analysis.  The  relative  amount  of  pure 
gum  and  of  cheaper  composition  may  vary,  or  good  material  may  be 
injured  by  defective  vulcanization.  The  only  safe  way  to  secure  excel- 
lence and  uniformity  is  for  the  pump  manufacturer  to  test  samples 
of  each  new  lot  under  severe  duty  (as  by  a  week's  run  in  a  small  special 
pump,  with  say  150  pounds  pressure  and  heavy  water  hammer,  or  by 
some  equivalent  means)  and  to  furthermore  require  the  rubber  manu- 
facturer to  mould  a  date  mark  as  "  (Name  of  pump  manufacturer, 
lot  201 — April  3,  1904)  "  on  the  edge  of  every  valve,  by  which  the  pump 
manufacturer  can  keep  track  of  those  which  prove  defective. 

32.  Size  and  Number  of  Pump-Valves,     a.  The   diameter  of  the   disc 
of  rubber  forming  the  valve  must  not  be  greater  than  4  inches  or  less  than 
3  inches. 

Three  and  five-eighths  inches  diameter  appears  to  be  the  size  best 
*neeting  all  the  conditions,  and  has  been  adopted  by  several  manufac- 
turers but  is  not  insisted  upon. 


SPECIAL   PUMPING   MACHINERY 


481 


There  is  some  confusion  between  different  shops  about  designating 
size  of  valves.  The  practice  is  here  adopted,  which  is  much  the  most 
widely  used,  of  naming  the  diameter  of  the  disc  of  rubber  which  covers 
the  ports,  and  it  is  hereby  specified  that  this  shall  be  about  one-half 
inch  greater  than  the  diameter  of  the  valve-port  circle  which  it  covers, 
thus  affording  about  one-fourth  inch  over-lap  or  bearing  for  the  rubber 
disc  all  around  its  edge. 

If  valves  are  larger  than  four-inch  there  is  an  increased  tendency 
to  valve-slam  at  the  very  high  speed  at  which  the  pump  is  designed 
to  run,  and  if  valves  are  smaller  than  three  inches  diameter  the  greater 
number  tends  to  unnecessary  multiplication  of  parts,  and  the  ports 
being  so  small  are  a  little  more  liable  to  become  obstructed  by  rubbish. 

b.  The  thickness  of  the  rubber  valve  must  in  no  cases  be  less  than  f -inch. 

33.  Suction  Valve  Area.  a.  The  total  lift  of  suction  valves  must  not 
exceed  |-inch. 

b.  The  net  suction  valve  port  area  and  the  total  suction  valve  outlet  area 
under  valves  lifted  J  inch  high  must  not  be  smaller  than  the  figures  given  in 
the  table  below: 


Approx.  Actual 

(6) 

Net 

Max.  Piston 

Suction 

(7) 

(l) 

Length 
of  Stroke 
(in 
inches)  . 

(2) 

Greatest 
No. 
Revolu- 
tions per 
Minute. 

Corre- 
sponding 
Piston 
Travel  per 
Minute. 

Velocity  at  Full 
Speed  per  Column 

(3)X2.2. 

Valve-port 
Area  Re- 
garded 
Necessary 
for  this 
Speed 
Per  Cent 
of  Plunger 

Suction 
Valve-out- 
let Area 
Under 
Valves 
Lifted  * 
Inch  High. 

& 

charge 
Valve 
Area. 

(4) 
Feet 

(s) 

Feet 

per  Min. 

per  Sec. 

Area. 

12 

70 

140  ft. 

308 

5-i 

56% 

56% 

1  of 

Suction 

16 

60 

1  60  ft. 

352 

5-9 

64% 

64% 

Valve 
Area. 

By  "  valve-outlet  area,"  we  mean  the  vertical  cylindrical  surface 
over  the  outer  edge  of  the  valve  ports,  i.e.,  the  distance  L  multiplied 
by  the  circumference  at  the  outer  edge  of  the  valve  ports  C.  Thus  for 
a  four-inch  valve,  with  ports  inscribed  in  a  three  and  one-half-inch 
circle,  whose  circumference  is  3.5x3.1416  =  11  inches;  the  valve  "out- 
let area"  for  one-half-inch  lift  would  be  5^  inches. 

The  actual  velocity  of  piston  during  the  middle  portion  of  stroke 
is  from  2.0  to  2.4  (average  2.2)  times  as  great  as  the  piston  travel  per 
minute  (as  determined  in  experiments  by  Mr.  J.  R.  Freeman  on  several 
duplex  pumps  of  different  manufacture).  This  is  because  each  piston 
stands  still  nearly  half  the  time,  or  while  its  mate  is  working,  and,  more- 
over, moves  more  slowly  near  start  and  finish  of  stroke.  The  words 
"piston  speed "  are  commonly  incorrectly  used,  and  refer  to  "piston 
travel."  A  clear  understanding  that  the  actual  piston  speed  is  more 
than  twice  as  great  leads  to  more  generous  valve  design. 

Large  aggregate  valve  areas  are  necessary  for  pumps  designed  to  run 
as  fast  as  these,  and  experience  has  shown  that  to  prevent  valve  slam 
at  high  speed  and  to  accommodate  high  suction  lifts,  it  is  just  as  important 
to  have  a  large  "  valve  outlet  area  "  as  to  have  a  large  area  of  valve  port. 

It  is  valve  slam  or  water  hammer  which  commonly  limits  the  highest 
speed  at  which  a  pump  can  be  run.  This  water  hammer  may  originate 
from  the  pulsations  in  a  long  or  small  suction  pipe.  The  vacuum  chamber 


482 


PUMPING  MACHINERY 


lessens  it,  but  there  is  commonly  some  point  of  high  water  in  the  vacuum 
chamber  that  will  give  much  smoother  action  than  any  other. 

Valve  slam  in  this  style  of  pump  is  caused  chiefly  by  the  short  rebound 
of  the  steam  piston  against  the  elastic  steam  cushion  at  the  end  of  the 

stroke.  This  in  turn  snaps  the  valves 
down  with  a  jump  when  the  speed  is  high. 
Dividing  this  impact  or  slam  on  numerous 
valves  of  low  lift,  tends  to  break  up  and 
lessen  the  shock,  therefore  with  valves  of 
the  size  and  style  used  in  fire  pumps,  other 
things  being  equal,  the  less  they  have  to 
rise  and  drop  to  let  the  water  through 
them,  the  less  will  be  the  valve  slam.  This 
height  of  rise  and  drop  is  governed  by  the 
circumference  rather  than  the  port  area. 
Experience  and  practice  have  shown  that  a 
£ -inch  limit  of  lift  is  reasonable  and  does 
insure  a  smooth  working  pump  under  all 
ordinary  conditions. 

c.  The  following  table  gives  minimums  for 


FIG.  377. — Water  Valve. 


aggregate  valve  port  area  and  aggregate  valve  outlet  area  for  the  different 
size  plungers,  figured  on  a  basis  of  56  per  cent  of  plunger  area  for  a  1 2-inch 
stroke,  and  64  per  cent  for  a  i6-mch  stroke. 


Size  of  Pump. 

500  gal. 

75o  gal. 

1000  gal. 

1500  gal. 

I 

Diameter  of  plunger.      Inches.  .  .  . 

7i" 

9" 

10" 

12" 

2 

Area  of  plunger  in  sq    inches 

41     28 

63   62 

78     S4 

1  T  7       TQ 

3 

56%  of  plunger  area,  or  minimum 
aggregate  valve-port  area  al- 
lowed per  section.  Sq.  inches.  . 

23.11 

35-63 

43-98 

64%- 
72.38 

4 

Minimum  aggregate  valve-port 
circumference,  allowed  per  sec- 
tion. Inches.  .  .  

46  .  22 

71  .  26 

8?     06 

144.  76 

5 

Minimum  aggregate  valve  outlet 
area  allowed  per  section  for 
valves  lifted  \  inch  high.  Sq. 
inches 

22       II 

•7  r    6  3 

A  -I      08 

72     -?8 

d.  If  we  consider  using  any  one  of  the  three  sizes  of  valves  below,  whose 
port  areas  may  be  assumed  approximately  as  given,  then  the  necessary  number 
of  valves  per  section  will  be  as  in  the  table  following: 


Diam.  Valve. 

Diam.  of  Valve- 
port  Circum 

Circum.  of  V.  C. 
Circle. 

Valve-port  Area  (Net). 
Square  Inches. 

3" 
3t" 

4" 

4" 

3t" 
3  1" 

7-8S" 
9.82" 
10.99" 

3-5 

5-1 

6.3 

SPECIAL  PUMPING   MACHINERY 


4S3 


Size  of  Pump. 

500  gal 

750  gal. 

1000  gal. 

1500  gal. 

Size  of  valves,  inches  

3 

3i. 

4 

3. 

3l 

4 

3 

3f 

9 

4 
8 

3 
19 

21 

3f 
14 
*5 

4 
14 

12 

Necessary  number  of  valves 
to  satisfy  (4)  under  c  

6 

S 

6 

9 

8 

7 

ii 

Necessary  number  of  valves 
to  satisfy  (3)  under  c. 

7 

5 

4 

10 

8 

6 

*3 

9 

7 

The  exact  number  and  size  of  valves  will,  however,  not  be  insisted  upon 
provided  the  aggregate  valve  area  and  the  aggregate  valve  outlet  area  for 
each  section  is  not  less  than  that  given  in  the  table  under  c  for  the  limiting 
lift  of  \  inch. 

Manufacturers  will  note  that  with  the  established  lift  of  £-inch, 
the  3 f -inch  valve  will  permit  a  valve  outlet  area  more  nearly  equal 
to  its  port  area  than  will  either  the  3 -inch  or  4-inch  valves,  and  a 
relatively  less  number  of  valves  will  satisfy  the  specifications. 

34.  Delivery  Valves,     a.  The  total  lift  of  delivery  valves  must  not  exceed 
\  inch. 

This  is  to  avoid  valve  slam,  as  explained  in  Art.  33. 

b.  The  aggregate  valve-port  area  should  be  restricted  to  about  two-thirds 
the  suction-valve  area. 

A  small  restriction  of  water-way  through  the  delivery  valves  steadies 
the  action  of  the  pump  and  tends  to  prevent  undue  pulsations  of  pres- 
sure in  the  delivery  pipe  or  fire  hose.  Fewer  delivery  valves  than  suction 
valves  are,  therefore,  preferred,  and  if  extra  holes  in  the  delivery  deck 
are  cast  for  shop  purposes  these  had  better  be  plugged  than  fitted  with 
valves. 

The  suction  valves  require  more  generous  port-circumference  and 
port-area  than  delivery  valves  because  when  a  pump  has  to  suck  its 
supply  through  a  considerable  height  or  through  a  long  pipe  there  should 
be  the  least  practicable  waste  of  the  atmospheric  pressure  in  getting 
the  water  into  the  plunger  chamber,  or  in  retarding  it  from  following  the 
plunger  in  full  contact.  With  the  water  once  in  the  plunger  chamber 
there  is  plenty  of  steam  pressure  available  to  force  it  out  through  the 
delivery  valves. 

35.  Valve  Springs,   Guards  and  Covers,     a.  All  valve  springs  must  be 
of  the  best  spring  brass  wire,  and  must  be  coiled  on  a  cylindrical  arbor. 

Conical  valve  springs  are  not  approved  because  the  stress  is  not 
uniform  throughout  spring,  thereby  increasing  the  liability  to  breakage 
and  the  chance  of  their  getting  out  of  center  and  becoming  "  hooked  up." 

b.  The  valve  spring  must  be  held  centrally  at  its  top  by  resting  in  a  groove 
in  valve  guard,  substantially  as  shown  in  Fig.  378. 

c.  A  light,  rustless, metallic  plate  must  be  interposed  between  the  bottom 
of  the  spring  and  the  rubber  valve,  and  must  be  the  full  area  of  the  valve. 
This  plate  must  also  be  formed  with  a  raised  bead  to  guide  the  spring  at  the 
bottom. 

The  weight  of  this  plate  should  be  small,  for  the  inertia  of  the  lifting 


484 


PUMPING  MACHINERY 


parts  of  the  valves  should  be  the  least  possible,  to  permit  quick  action 
and  to  avoid  pounding. 

d.  For  the  average  condition  of  a  10-  or  15 -foot  lift,  the  stiffness  of  suction 
valve  springs  should  be  such  that  a  force  of  about  one  pound  per  square  inch 
of  net  port  area  will  lift  valve  J-inch  off  its  seat. 

The  springs  on  the  delivery  valves  should  ordinarily  be  from  two  to  three 
times  as  stiff  as  just  specified,  but  any  other  reasonable  degree  of  stiffness 
which  is  proved  to  work  well  in  practice  will  not  be  objected  to. 

For  suction  under  a  head,  the  greater  snap  with  which  water  enters 
the  plunger  chamber  when  thus  pushed  in  by  say  twice  the  atmospheric 
pressure  renders  it  difficult  to  avoid  water  hammer 
at  high  speed.  Extra  stiff  suction  valve-springs 
will  commonly  aid  in  controlling  this  and  should 
be  used  wherever  pumps  are  to  work  under  a 
head. 

An  approved  type  of  indicator  water  gate  on 
the  suction  pipe  near  the  pump,  which  can  be 
partly  closed,  will  enable  the  pump  to  run  quietly 
at  high  speed.  Such  a  gate  is  an  extra  not  in- 
cluded in  price  of  the  pump. 

36.  Sticking    of    Valves,     a.   Steam    fire-pumps 
should  be  started  to  limber  them  up  at  least  once  a 
week. 

Although  vulcanized  India  rubber  is  much 
the  best  material  yet  used  for  fire-pump  valves, 
unfortunately  the  brass  is  sometimes  corroded  by 
the  free  sulphur  contained  in  the  rubber,  so  that 
if  the  pump  is  left  standing  for  several  weeks  the 
rubber  valve  discs  may  become  stuck  to  their 
brass  seats,  and,  if  suction  has  a  high  lift,  there 
may  not  be  vacuum  enough  to  tear  all  the  suction 
valves  open  when  pump  is  started. 

37.  Valve  Seats,     a.  All  water  valve  seats  muct 
be    of    bronze    composition.     They   may   be    either 


FIG.  378. — Valve. 


screwed  into  the  deck  on  a  taper  or  forced  in  on  a  smooth  taper  fit.  With 
either  arrangement,  the  seat  must  be  either  flanged  out  on  the  under  side  all 
the  way  round  or  be  provided  with  a  substantial  lug  opposite  each  rib,  these 
lugs  being  expanded  out  after  the  valve  is  inserted. 

If  the  valve  seats  are  not  expanded  after  being  put  in  place,  there 
is  a  possibility  that  now  and  then  a  valve  seat  will  work  loose  and  come 
out,  thus  crippling  the  pump. 

b.  The  under  side  of  the  valve  deck  must  be  rounded  over  to  give  good 
bearing  for  the  expanded  part  of  the  seat. 

c.  Three-inch  valves  must  have  four  or  five  ribs,  3^-inch  valves  five  or 
six  ribs,  and  4-inch  valves  six  ribs. 

Enough  ribs  must  be  provided  to  give  proper  support  to  the  rubber 
valve,  but  too  many  are  objectionable,  as  small  ports  would  be  liable 
to  obstruction  by  refuse. 

d.  The  edges  of  the  valve-seat  ports  must  be  moderately  rounded  over 
to  remove  such  sharp  edges  and  points  as  would  be  liable  to  cut  or  damage 
the  rubber  valve  when  under  pressure. 


SPECIAL  PUMPING   MACHINERY 


485 


38.  Valve  Stems,  a.  All  valve  stems  must  be  of  f-inch  Tobin  bronze 
and  of  the  fixed  type,  and  must  have  the  guard  fastened  on  by  one  of  the 
methods  shown  by  Figs.  378  and  379. 

Other  methods  may  be  approved,  in  writing,  if  found  by  test  and 
experience  to  have  especial  merit. 

b.  These  stems  must  be  screwed  into  the  seats  on  a  straight,  tightly  fitting 
thread,  and  the  lower  end    then  well    headed  over 
into  a  countersink.      The  valve  guard  and  nut  must 
be  of  composition. 

In  Fig.  378  the  upper  part  of  the  stem  is  slab- 
bed off  on  two  opposite  sides  and  fits  a  correspond- 
ing hole  in  the  guard. 

The  guard,  therefore,  cannot  turn.  The  out- 
side of  the  special  nut  is  fitted  on  a  taper  to  the 
inside  of  the  guard,  and  the  nut  tapped  out  to  fit 
the  five-eighths  U.  S.  thread  on  the  stem. 

The  action  of  the  valve,  whether  with  the  spring 
or  without,  tends  to  drive  these  taper  fits  together, 
producing  a  frictional  lock  similar  to  that  of  a 
friction  clutch ;  and  although  the  nut  may  be  loose 
on  the  thread,  it  cannot  possibly  work  off. 

It  will  be  apparent  that  the  taper  fit  on  the 
nut  must  be  so  made  as  to  always  bear  on  the 
taper  fit  in  the  guard,  and  not  bottom  in  the  guard. 

It  is  believed  that  with  the  present  screw 
machine  practice  in  shops  of  to-day  these  small 
parts  can  readily  be  turned  out  accurately  and 
cheaply  in  large  quantities.  The  nuts  and  guards 
made  in  any  one  shop  must  be  exactly  of  stand- 
ard dimensions,  so  that  the  product  of  different 
periods  will  be  interchangeable. 


FIG.  370. — Valve. 


The  taper  should  be  about  one  inch  to  one  foot.  With  this  taper 
the  nut  can  be  readily  turned  in  or  out,  but  there  is  friction  enough  to 
hold  the  guard  and  nut  together  even  if  the  spring  is  off. 

In  Fig.  379  the  top  of  the  guard  is  recessed  in  the  form  of  a  hollow 
inverted  pyramid  of  six  sides,  to  correspond  to  a  hexagonal  nut.  The 
angle  of  two  opposite  sides  of  this  recess,  which  should  be  about  75  degrees, 
will  both  surely  lock  the  nut  and  still  permit  of  its  being  turned  with  a 
wrench. 

The  guard  is  kept  from  turning  by  slabbing  off  the  stem  in  the  same 
manner  as  described  and  shown  in  Fig.  378. 

To  facilitate  the  removal  of  the  nut,  the  edges  should  be  slightly 
chamfered.  An  unfinished  nut  simply  drilled  and  tapped  is  all  that  is 
desired.  Any  hexagonal  or  square  nut  within  the  size  of  the  tapered 
recess  will  be'  locked. 

With  this  construction,  the  nut  cannot  turn  in  either  direction  without 
compressing  the  spring  and  is  therefore  locked,  and,  in  the  event  of 
the  spring  breaking  or  being  left  off,  the  nut  is  well  protected  in  its 
recess,  from  the  possible  turning  effects  of  water  currents,  and  experi- 
ments have  shown  that  it  will  stay  in  place. 

With  machine  molding  it  will  be  possible  to  make  these  guards 
complete  in  foundry,  requiring  no  machine  work  further  than  a  pos- 
sible broaching  out  of  hole  to  fit  the  stem,  as  a  fairly  good  fit  is 
necessary. 

While  both  of  these  devices  are  effective  even  though  not  tightened 
down  to  a  shoulder,  they  should  be  so  tightened  for  greater  safety  and 
to  fix  the  lift  at  the  half -inch  limit.  - 


PUMPING  MACHINERY 


39.  Pipe  Sizes.  .  a.  Water  and  steam  pipe  connections  must  have  standard 
flanges  to  connect  with  pipes  of  the  sizes  given  below. 


1 

• 

Size  of  Pump. 
Gal.  per  Min. 

Diameter  of 
Suction  Pipe. 
Inches. 

•Diameter  of 
Discharge  Pipe. 
Inches. 

Steam 
Pipe. 

Exhaust 
Pipe. 

500 

8 

6 

3 

4 

750 

10 

7  or  8* 

Si 

4 

IOOO 

12 

8 

4 

5 

1500 

14 

10 

5 

6 

*Eight-inch  preferred,  this  being  the  more  common  size  for  valves,  fittings  and  pipes. 

These  suction  pipe-sizes,  although  larger  than  common  for  trade 
pumps  of  the  same  size,  are  believed  to  be  amply  justified  by  experience, 
and  exert  a  powerful  influence  toward  enabling  the  pump' to  run  smoothly 
at  high  speed  with  water  cylinders  filling  perfectly  at  each  stroke.  No 
defect  is  more  common  than  restricted  suction  pipes. 

b.  A  single  suction  entrance  at  the  end  of  the  pump  is  to  be  provided 
unless  otherwise  specified  by  the  purchaser. 

Some  situations  render  desirable  side  suction  entrances,  for  permitting 
drafting  water  from  two  different  sources  of  supply.  These  additional 
openings  are  to  be  considered  as  extras.  Ordinarily,  the  purchaser 
can  provide  for  such  situations  by  proper  piping  at  the  single  end  suction 
entrance. 

If  there  is  to  be  but  one  suction  opening  on  casting,  this  had  best 
be  at  center,  for  the  reason  that,  if  suction  pipe  ever  gets  to  leaking  air, 
this  air  stands  a  better  chance  of  being  distributed  equally  to  the  two 
plungers,  and  has  less  tendency  to  make  the  pump  run  unevenly. 

c.  Standard  flanges  and  standard  bolt  layouts  as  adopted  by  the  Master 
Steam  Fitters,  July  18,  1894,  must  be  used  on  all  the  above  pipe  connections, 
as  per  table  given  below. 

SCHEDULE  OF  STANDARD  FLANGES 


Size  of  Pipe 
X   Diam. 
of  Flange. 
Inches. 

Diameter 
of  Bolt  Circle. 

Inches. 

Number 
of  Bolts. 

Size  of 
Bolts. 

Inches. 

Flange 
Thickness 
at  Edge. 
Inches. 

3    X    7i 

6 

4 

IXrt 

»'. 

3*X    8} 

7 

4 

IXai 

I 

4X9 

7l 

4 

|Xt| 

if 

4iX    9t 

7l 

8 

1X3 

it 

5    Xio 

H 

8 

|X3 

if 

6    Xn 

9* 

8 

|X3 

I 

7    Xiai 

ioi 

8 

iX3f 

*& 

8   Xi3i 

iif 

8 

1X3* 

i| 

9    Xis 

•ji 

12 

iX3l 

il 

10    Xi6 

14} 

12 

lX3f 

'A 

12    X  19 

J7 

12 

|X3i 

ii 

14    X2i 

i8i 

12 

i    X4i 

if 

SPECIAL   PUMPING   MACHINERY  487 

Do  not  drill  bolt  holes  on  center  line,  but  symmetrically  each  side  of  it. 
On  steam  and  exhaust  openings  loose  flanges  threaded  for  wrought-iron 
pipe  must  be  provided. 

Where  the  situation  will  not  permit  of  a  standard  flange  on  exhaust 
opening  for  lack  of  room,  a  special  flange  threaded  to  fit  the  proper  size 
wrought-iron  pipe  may  be  used. 

40.  Air  and  Vacuum  Chambers,  a.  Air  and  vacuum  chambers  in  accord- 
ance with  the  sizes  given  in~  the  following  table  must  be  provided  with  all 
pumps.  If  the  air  chamber  is  cast  iron,  the  pump  manufacturers  must  warrant 
that  it  has  been  subjected  to  a  hydraulic  test  of  400  pounds  per  square  inch 
before  it  is  connected  to  pump. 

It  is  to  be  thoroughly  painted  inside  and  out  to  diminish  its  porosity. 

SIZE   OF  VACUUM    AND  AIR  CHAMBERS 


Vacuum  Chamber 
is  to  Contain 

Air  Chamber 
is  to  Contain 

coo  gallon  pump 

13  gallons. 

1  7  gallons. 

7  co        '  ' 

18 

25        '  ' 

24 

30         '  ' 

I  ZOO         '  ' 

•?o 

40 

i^WVJ 

The  air  chamber,  combined  with  connections  for  discharge  pipe, 
relief  valve,  and  hose  valves,  should  be  carefully  designed  to  make 
the  whole  weight  as  small  as  possible.  Keeping  this  weight  down  makes 
the  pump  run  steadier  and  brings  less  stress  on  the  flanges  at  high 
speeds. 

An  air  chamber  of  hammered  copper  and  warranted  tested  under  a 
hydraulic  pressure  not  less  than  300  pounds  per  square  inch  is  a  little 
better  than  cast  iron,  as  it  holds  air  better,  and  being  lighter  it  wrenches 
and  strains  the  pump  less  when  running  fast  and  shaking,  but  because  it 
costs  from  $25  to  $50  more  than  cast  iron,  it  is  not  often  adopted. 

b.  The  vacuum  chamber  must  be  attached  to  the  pump  in  the  most  direct 
way  practicable,  but  provision  must  be  made  for  attaching  it  in  such  manner 
as  not  to  prevent  readily  taking  off  the  cylinder  heads. 

c.  Every  vacuum  chamber  should  be  provided  on  one  side  near  the  top 
with  a  |-inch  pipe  tap  plugged.     This  to  be  used  for  attaching  a  vacuum 
gauge  if  desired. 

41'.  Pressure  Gauge,  a.  A  pressure  gauge  of  the  Lane  double  tube  spring 
pattern  with  5-inch  case,  must  be  provided  with  the  pump,  and  connected 
near  to  inboard  side  of  air  chamber,  as  shown  in  Fig.  381,  by  a  J-inch  cock, 
with  lever  handle. 

The  dial  of  this  gauge  should  be  scaled  to  indicate  pressures  up  to  240 
pounds,  and  be  marked  " Water." 

This  kind  of  gauge  is  used  on  locomotives  and  is  the  best  for  with- 
standing the  vibration  which  causes  fire-pump  gauges  to  be  often  unre- 
liable, Moreover,  this  double  spring  form  is  safer  against  freezing. 


488 


PUMPING  MACHINERY 


42.  Hose  Valves,     a.  Hose  valves  must  be  attached  to  the  pump  (and 
included  in  its  price)  as  follows: 

For  the  2  stream  or  5oo-gal.  pump,  2  hose  valves. 

For  the  3  stream  or  75o-gal.  pump,  3  hose  valves. 

For  the  4  stream  or  looo-gal.  pump,  4  hose  valves. 

For  the  6  stream  or  i5oo-gal.  pump,  6  hose  valves. 

These  are  to  be  2^-inch  straightaway  brass  valves,  without  cap, 
and  similar  and  equal  in  quality  to  those  made  by  the  Chapman  Valve 
Company,  the  Ludlow  Valve  Company,  or  the  Lunkenheimer  Company. 

The  hose-screw  at  end  of  these  valves  is  to  be  fitted  to  a  hose  coupling 
furnished  by  the  customer,  or  where  this  cannot  be  procured  may  be 
left  with  the  thread  uncut. 

To  accommodate  locations  where  all  the  lines  of  hose  must  lead  off 
from  one  side  of  the  pump — makers  can  furnish  a  spool  piece  or  special 
casting  to  which  the  hose  valves  can  be  attached — but  this  is  an  extra 
not  included  in  the  regular  price. 

43.  Safety  Valve,     a.  A  safety  or  relief  valve  of  the  Ashton,   Crosby, 
American,  or  other  make  agreed  upon  in  writing  with  this  office,  is  to  be  regu- 
larly included  in  the  price,  and  is  to  be  attached  to  each 
pump;    preferably  extending   horizontally   inboard   from 
base  of  air  chamber,  as  shown  in  Fig.  381,  so  that  its 
hand-wheel  for  regulating  pressure  is  within  easy  reach. 
This  hand-wheel  must  be  marked  very  conspicuously,  as 
shown  in  sketch. 

b.  This  valve  is  to  be  set  ordinarily  at  a  working 
pressure  of  100  pounds  to  the  square  inch,  and  is  to  be  of 
such  capacity  that 
when  set  at  100 

pounds  it  can  pass  all  the  water  dis- 
charged by  the  pump   at  full  speed, 
at  a  pump  pressure  not  exceeding  125 
pounds  per  square  inch. 
For    5oo-gallon  pump,  a  3-inch  valve. 
For    75o-gallon  pump,  a  3^-inch  valve. 
For  icoo-gallon  pump,  a  4-inch  valve. 
For  i5oo-gallon  pump,  a  5 -inch  valve. 
The  relief  valve  must  discharge  in 
a  vertical  downward  direction  'into  a 
cone  or  funnel  secured  to  the  outlet  of 
the  valve.     (See  Art.  44.) 

The  valve  must  be  so  attached  to 


FIG.  380. — Hand 
Wheel. 


FIG.  381.— Safety  Valve.  ' 


the  delivery  elbow  and  discharge  cone  by  flange  connections  as  to  permit  of 
its  ready  removal  for  repairs  without  disturbing  the  waste  piping. 

44.  Discharge  Cone.  a.  This  cone  should  be  so  constructed  that  the 
pump  operator  can  easily  see  any  water  wasting  through  the  relief  valve, 
and  its  passages  should  be  of  such  design  and  size  as  to  avoid  splashing 
water  over  into  the  pump  room. 

b.  The  cone  must  be  provided  with  an  opening  to  receive  the  air  vent  pipe 


SPECIAL   PUMPING  MACHINERY 


489 


required  by  Art.  45,  and  the  arrangement  must  be  such  that  the  pump  operator 
can  easily  tell  whether  water  is  coming  from  the  air  pipe  or  is  wasting  through 
the  relief  valve. 

c.  The  cone  should  be  piped  to  some  point  outside  of  the  pump  house 
where  water  can  be  wasted  freely,  the  waste  pipes  being  as  below. 


Size  of  Pump. 

Diameter  of  Waste 
Pipe  from  Cone. 

500  gallon  
75°               

IOOO                     
I  <OO         '  ' 

5  inches. 
6       " 

7 
8 

The  waste  pipe  can  pass  down  to  floor  between  the  yokes  at  middle 
of  pump.  It  should  be  piped  in  such  a  way  that  steam  and  gases  from 
other  drains  or  waste  pipes  will  not  work  back  through  it,  and,  by  being 
troublesome  in  the  pump  room,  suggest  the  covering  of  the  cone  in  any 
way,  as  it  is  desirable  that  the  pump  operator  should  always  be  able  to 
see  instantly  any  waste  from  the  relief  valve  or  air  vent. 

This  cast-iron  cone,  connected  to  the  safety  valve  and  air  vent, 
is  included  in  price  of  pump,  but  the  waste  pipe  beyond  it  is  not. 

45.  Air  Valve,     a.  An  air  vent  with  a  brass  gate  valve  and  brass  pipe  for 
connecting  up  must  be  provided  and  connected  with  delivery  elbow  and 
discharge  cone. 

b.  The  size  of  this  air  vent  should  be  i  inch  for  5oo-gallon  and  75o-gallon 
pumps,  and  i\  inches  for  the  looo-gallon  and  i5oo-gallon  sizes. 

c.  The  hand-wheel  of  this  valve  must  be  marked  as  per  sketch.     The 
lettering  must  be  very  open,  clear,  and  distinct,  not 

liable  to  be  obstructed  by  grease  and  dirt,  and  of  a 
permanent  character. 

The  object  of  this  valve  is  to  reduce  the  pres- 
sure above  force  valves  and  secure  a  prompt  riddance 
of  all  air  that  may  come  through  the  water  cylinders 
when  first  starting  up. 

This  valve,  of  course,  should  be  closed,  when 
once  pump  is  under  way,  to  prevent  waste  of  water. 

46.  Priming,     a.  Each  pump  must  be  fitted  with 
a  set  of  brass  priming  pipes  and  valves,  according  to 
either  one  or  the  other  of  the  following  methods : 

b.  For  looo  and  i5oo-gallon  pumps,  the  priming  pipes  must  be  i|  inch. 
For  the  500  and  75o-gallon  pumps,  the  pipes  must  be  i  inch.  Pump-makers 
are  to  furnish  these  pipes  and  the  fittings  called  for  below,  and  are  to  connect 
them  up  providing  a  2-inch  outlet,  looking  upwards,  ready  for  the  supply 
from  the  priming  tank. 

The  pipe  from  the  priming  tank  to  this  outlet  should  be  at  least 
two-inch,  and  may  be  of  iron,  and  is  to  be  furnished  by  the  purchaser. 
All  parts  furnished  by  the  pump-maker  are  to  be  of  brass,  and  are  to  be 
included  in  the  price  of  the  pump. 

Controllable  Valve  Arrangement,  c.  Four  two-seat  controllable  valves,  one 
for  each  pulsation  chamber,  and  of  the  general  type  illustrated  in  Fig.  383, 


FIG.  382.— Valve 
Handle. 


490 


PUMPING   MACHINERY 


must  be  provided.    In  these  the  inlet  of  water  and  the  outlet  of  air  are  simul- 
taneously opened  and  closed  by  the  pump  operator. 

This  valve  can  preferably  be  provided  with  a  flange  connection  in 
place  of  the  threaded  one,  and  secured  to  water  cylinder  with  three 
five-eighth  bolts.  This  will  permit  of  easier  fitting  up  as  to  pipe  con- 


FIG.  383.— Valves. 

nections.  Objection  has  been  raised  to  this  double-seated  valve  from 
the  possible  difficulty  of  keeping  both  seats  tight.  If  desired,  the  stem 
between  the  two  seats  may  be  somewhat  enlarged  and  provided  with  a 
suitable  spring,  thus  giving  flexibility  between  the  two  seats  and  pre- 
venting all  trouble  from  uneven  wear. 

d.  The  hand-wheel  of  each  of  these  valves  must  be  marked  as  per  sketch 
so   that   the  pump  operator   may   clearly  understand 
their  use.    The  lettering  must  be  very  open,  clear,  and 
distinct,  not  liable  to  be  obscured  by  grease  and  dirt, 
and  of  a  permanent  character. 
e.  There  must  be  provided 
and  fitted  to  each  combined 
valve  a  check  and  umbrella- 
top  air  vent,  as  shown  in  Fig. 
385.    This  fitting  must  have  a 
clear  passageway  through  it, 
the  full  equivalent  of  a  half -inch  bore. 
The  check  valve  is  to  permit  the  outflow  of  air,  but  to  prevent  the 
influx  when  the  plunger  is  sucking. 


FIG.  ^84—  Valve 
Handle. 


FIG.  385.— Check  Valve. 


SPECIAL   PUMPING   MACHINERY  491 

This  method  is  preferred  to  the  one  using  rubber  priming  checks, 
as  now  and  then  a  rubber  valve  will  stick  on  its  seat  and  thus  prevent 
priming  of  one  of  the  chambers.  In  this  arrangement  the  pump  operator 
has  absolute  control  over  the  priming  water  into  each  chamber. 

Another  advantage  is  that  the  connection  of  the  air  vent  with  the 
priming  valve  ensures  that  the  air  vents  will  be  opened ;  and  further,  by 
the  vigorous  spurting  out  of  water  as  soon  as  the  pump  is  primed,  the 
pump  operator  is  reminded  that  the  priming  valve  should  be  closed. 

Should  the  pump  operator,  however,  through  a  mistaken  idea  of  the 
proper  method  of  operation,  think  that  the  priming  should  be  con- 
tinued until  all  air  was  exhausted  from  the  suction  pipe,  and  the  pump 
running  in  normal  condition,  there  would  be  some  by-passing  between 
chambers,  but  as  there  is  a  free  vent  for  the  air,  the  main  result  would 
be  simply  to  limit  the  amount  of  air  exhausted  per  stroke,  from  the  main 
suction,  by  the  amount  of  water  which  entered  a  chamber  in  this  way. 
The  amount  of  water  thus  entering,  however,  would  not  be  appreciably 
greater  than  that  which  would  enter  from  the  priming  tank  with  the 
check- valve  arrangement. 

If,  in  spite  of  the  warning  given  by  the  spurting  air  vents,  the  pump 
operator  should  neglect  to  close  the  priming  valves  when  the  pump  was 
running  normally,  the  priming  tank  would  eventually  be  overflowed; 
but  this  would  not  be  as  serious  as  the  drawing  in  of  air  from  an  exhausted 
priming  tank,  which  would  result  with  the  check-valve  method,  were  the 
main  two-inch  valve  similarly  neglected. 

Rubber  Check  Valves,  f.  Four  rubber  check  valves,  one  for  each  pulsation 
chamber,  and  similar  to  ordinary  pump  valves,  must  be  provided.  The 
chambers  for  these  should  preferably  be  made  as  a  part  of  the  pump  cylinder, 
thus  securing  a  compact  arrangement. 

g.  The  valve  seat  should  have  three  ribs  to  the  central  hub,  supporting 
the  rubber  valve.  The  net  port  area  through  the  valve  should  not  be  less 
than  one  and  one-half  square  inches. 

This  valve  seat  should-  rest  in  an  inverted  position,  and  can  be  so 
fitted  up  as  to  be  readily  removed.  The  valve 
stems  can  be  of  the  removable  type  screwing  into 
the  seat,  but  must  be  made  long  enough  to  receive 
a  check  nut  on  the  opposite  side  of  seat.  This 
will  effectually  lock  the  stem  in  place. 

h.  Care  must  be  taken  to  arrange  the  water  pas- 
sages through  and  about  these  priming  checks,  so  as 
to  avoid  all  air  pockets  and  so  as  to  reduce  to  a  mini- 
mum the  possibility  of  the  valves  becoming  choked 
up  by  refuse. 

i.  The  valve  seats,  stems  and  all  parts  must  be 
of  composition  and  of  strong  rugged  design,  so  fitted 
up  that  there  is  the  least  chance  for  the  rubber  valves 
to  stick,  and  with  all  parts  securely  put  together  the 
valves  must  be  readily  accessible.  • 

;'.  The  valve  springs  must  have  only  sufficient  strength  to  keep  the  valves 
on  their  seats,  so  that  they  will  freely  open  even  with  the  low  head  of  priming 
water  often  existing. 

k.  There  must  be  provided,  and  attached  to  the  top  of  each  plunger  cham- 
ber, a  brass  check  valve  and  air  cock  with  umbrella  top,  as  shown  by  Fig. 


492  PUMPING  MACHINERY 

386.    This  cock  and  valve  must  have  a  clear  passageway  through  them — the 
full  equivalent  of  a  half -inch  bore. 

The  check  valve  is  to  permit  the  outflow  of  air,  but  to  prevent  the 
influx  when  the  plunger  is  sucking.  Cocks  with  lever  handles,  are  used, 
as  these  show  clearly  whether  they  are  open  or  shut. 

/.  There  must  also  be  provided  a  two-inch  brass  gate  valve  for  the  general 
control  of  the  water  to  the  four  check  valves.  The  hand-wheel  of  this  valve 
must  be  marked  as  per  Fig.  384.  The  lettering  must  be  very  clear,  open,  and 
distinct,  not  liable  to  be  obscured  by  grease,  and  of  a  permanent  character. 

It  is  essential  for  a  properly  working  pump  that  the  main  two-inch 
priming  valve  should  be  closed  as  soon  as  the  pump  is  primed.  Other- 
wise water  will  be  drawn  from  the  priming  tank,  lessening  the  lifting 
power  of  the  pump  through  the  main  suction,  and  if  this  is  continued 
the  priming  tank  will  often  be  exhausted  and  air  drawn  into  the  pump, 
interfering  with  its  proper  action.  It  is  for  this  reason  that  the  marking 
on  the  priming  valve  is  required. 

For  all  average  situations,  either  method  of  priming  permits  of  getting 
the  pump  under  way  in  a  very  few  minutes,  but,  for  cases  where  the 
suction  pipe  is  over  300  or  400  feet  in  length,  or  sometimes  where  the 
lift  is  over  18  feet,  or  where  there  is  a  combination  of  long  length  and 
lift  within  these  limits,  so  much  time  is  consumed  in  exhausting  the  air 
from  the  suction  pipe  that  it  becomes  desirable  to  supplement  this 
method. 

For  such  situations,  a  steam  ejector  connected  to  the  suction  pipe 
near  the  pump  is  advised,  and  may  be  required  in  addition  to  the  regular 
priming  pipes  and  tank.  The  size  of  the  ejector  should  be  sufficient 
to  exhaust  the  suction  pipe  within  about  three  minutes.  Such  ejectors 
will  be  considered  as  extras  not  included  in  the  ordinary  pump  fittings. 

For  cases  where  pump  can  only  take  its  suction  under  a  head,  if 
absolutely  certain  that  the  level  of  the  suction  water  will  never  fall 
below  level  of  center  of  pump,  these  priming  pipes  may  be  omitted,  but 
openings  for  them,  into  the  pump  shell  must  be  provided  and  capped 
or  plugged. 

A  foot  valve  on  a  fire- pump  suction  is  not  advised  except  in  very  rare 
cases,  as  with  a  lift  of  18  feet  or  a  suction  pipe  500  feet  or  more  long. 
A  foot  valve  is  not  needed  when  there  is  a  good  efficient  set  of  priming 
arrangements,  as  described  above,  and  it  is  commonly  found  it  gives  a 
false  sense  of  security,  and  that  with  a  fire  pump  left  standing  several 
days  the  water  will  often  be  found  to  have  leaked  back,  so  that  it  is  no 
better  than  if  no  foot  valve  had  been  used. 

A  foot  valve  must  of  necessity  generally  be  located  where  it  is  inac- 
cessible for  quick  repairs,  and  as  they  grow  old,  foot  valves  are  often 
a  source  of  trouble.  Where  a  suction  pipe  is  exposed  even  slightly  to 
frost,  a  foot  valve  is  especially  objectionable. 

A  priming  tank  is  provided  by  the  purchaser  in  all  cases  where  there 
is  ever  to  be  any  lift  on  the  suction.  It  is  generally  advised  that  this 
tank  have  a  capacity  of  one-half  of  what  the  pump  can  throw  at  full 
speed  in  a  minute.  This  means  250  gallons  for  a  5oo-gallon  pump,  and 
500  gallons  for  a  icoo-gallon  pump,  etc.  It  is  the  intention  to  make 
the  pump  a  truly  "  independent  source  "  of  supply,  therefore  the  need 
of  a  special  priming  tank. 

Older  Priming  Arrangements.  The  form  of  priming  arrangements  here- 
tofore used,  with  metal  check  valves,  one  main  two-inch  priming  valve,  and 
one- inch  priming  pipes,  separate  controllable  air  cocks,  may  be  retained  on 
all  pumps  at  present  in  service,  and  will  be  considered  satisfactory,  if  kept  in 
good  order. 


SPECIAL   PUMPING  MACHINERY  493 

If  in  any  case  such  checks  give  trouble  the  priming  arrangement  may  be 
changed  and  valves  like  Fig.  383  or  rubber  checks,  as  described  in  sections 
/-/,  made  up  in  detachable  form,  may  be  put  on  if  desired,  where  the  con- 
nections on  the  pump  permit  them. 

Where  neither  method  is  desired  or  where  neither  is  feasible  the  faulty 
checks  may  be  replaced  by  a  special  type  such  as  are  now  made  for  this  use 
by  the  Locke  Regulator  Company,  of  Salem,  Mass.  These  are  one-inch 
check  valves,  adapted  to  use  a  small  disc  of  medium  hard  rubber,  similar  to  a 
pump  valve. 

These  fittings  are  very  near  the  dimensions  of  the  commercial  check 
valve,  so  that  with  sight  shortening  of  piping  connections  they  will  fit  into 
the  present  arrangements,  and  give  satisfaction. 

47.  Drain  Cocks,     a.  Five  brass  drain  cocks,  each  with  a  lever  handle 
and  of  half-inch  bore,  are  to  be  provided,  and  located  one  on  each  end  of 
each  water  cylinder,  and  one  above  the  upper  valve  deck. 

Care  should  be  taken  to  select  a  pattern  of  cock  whose  passageway 
is  the  practical  equivalent  of  a  half-inch  hole.  Some  patterns  of  half-inch 
commercial  cocks  although  threaded  for  half-inch  pipe  thread  have  but 
a  quarter-inch  hole  through  them.  Such  are  not  acceptable. 

TESTS  FOR  ACCEPTANCE. 

48.  Test  for  Smoothness  of  Action,     a.  Provide  outlets  for  the  water; 
start  the  pump  slowly,  gradually  open  steam-throttle  to  bring  the  pump  to 
full  speed.      The  pump  should  run  smoothly  at  the  rated  full  speed  of  70 
revolutions  per  minute  (or  60  revolutions  if  a  i5oo-gallon  pump)  with  full 
length  of  stroke,  and  meanwhile  maintain  a  water  pressure  of  100  pounds 
per  square  inch. 

If  the  hose  lines  are  short,  or  discharge  is  too  free,  partly  close  the  water 
outlet  valves,  thus  throwing  an  extra  back  pressure  on  the  pump  equivalent 
to  that  which  would  be  produced  through  a  greater  length  of  hose. 

During  this  trial  it  is  preferable  to  discharge  the  water  through 
lines  of  aj-inch  cotton  rubber-lined  hose,  preferably  each  150  feet  long, 
each  connected  directly  to  the  hose  outlets  on  the  pump,  and  each  line 
having  a  i^-inch  smooth  nozzle  at  its  outer  end.  Two  lines  should  .be 
connected  for  a  5oo-gallon  pump,  three  for  a  750,  and  so  on,  having 
as  many  lines  as  rating  of  pump  requires. 

A  hose  Iine"i5o  feet  long,  with. an  inside  surface  of  average  smoothness, 
and  .with  a  i^-inch  nozzle  attached,  will  require  about  80  pounds  pressure 
at  the  pump  to  discharge  250  gallons  per  minute,  and  the  nozzle  pressure 
will  be  about  45  pounds.  Therefore,  with  lines  attached  as  above,  q, 
pressure  at  the  pump  of  about  80  pounds  should  represent  a  discharge 
about  equal  to  the  rated  capacity  of  the  pump,  and  would  ordinarily 
correspond  with  the  rated  full  speed  revolutions. 

If  the  pump  runs  smoothly  under  these  conditions,  it  is  well  to  open 
the  throttle  somewhat  further,  and  bring  the  pressure  at  the  pump  up 
to  100  pounds.  This  will  give  a  discharge  of  about  280  gallons  per  stream, 
or  about  12  per  cent  in  excess  of  the  rated  capacity.  The  revolutions 
will,  of  course,  correspondingly  increase,  and  under  all  ordinary  conditions 
a  pump  should  run  smoothly  at  this  higher  capacity,  though  a  little 
more  vibration  and  pounding  would  be  expected  than  when  running 
simply  at  its  rated  speed. 


494 


PUMPING  MACHINERY 


After  cushion  valves  are  adjusted  there  should  be  no  noteworthy 
water  hammer  or  valve-slam.  Sometimes  valve-slam  is  not  the  fault 
of  the  pump,  but  arises  from  an  obstructed  suction  pipe.  It  is  objection- 
able to  doctor  water  hammer  in  a  pump  by  shifting  air  into  the  suction, 
as  this  cuts  down  the  efficiency  and  is  a  poor  expedient. 

The  quietness  of  that  part  of  the  hose  near  the  pump,  or  its  freedom 
from  rubbing  back  and  forth  crosswise  an  inch  or  more  with  each  pulsa- 
tion of  the  pump,  is  a  good  index  of  the  pump  maker's  skill  in  securing 
uniform  delivery.  Bad  pulsation  quickly  wears  holes  in  the  hose,  and 
to  reveal  this  is  the  object  of  testing  with  hose  connected  directly  to  the 
pump. 

49.  Test  of  the  Internal  Friction,  a.  This  is  shown  by  the  reading  of 
steam  gauge  compared  with  water  pressure  gauge  at  air  chamber. 

Tests  have  generally  run  about  as  follows,  for  pumps  running  at  full 
rated  speed: 


Size  Gallons 
per  Minute 
Capacity. 

Ratio  of 
Steam  Piston 
Area  to 
Water  Piston 
Area. 

Water 
Pressure  Ibs. 
per  Sq.in. 

Steam 
Pressure 
Theoretically 
Necessary, 
Disregarding 
Friction. 

Excess  of 
Steam  Pres- 
sure Needed 
to  Overcome 
Friction  Back 
Pressure,  etc. 

Actual 
Steam  Pres- 
sure Found 
Necessary 
at  the  Pump. 

500 

4  Times 

IOO 

25 

15 

40 

75° 

3 

IOO 

33 

12 

45 

IOOO 

3 

IOO 

33 

12 

45 

1500 

2f 

IOO 

36-5 

13-5 

5° 

b.  The  steam  pressure  needed  will  vary  slightly  with  the  freedom  of  the 
exhaust  pipe  and  with  the  tightness  of  the  packings,  etc.,  but  a  steam  pressure 
of  45  pounds  at  the  steam  chest  should  suffice  for  100  pounds  water  pressure 
on  pump  in  proper  adjustment. 

50.  Test  of  Strength  and  Tightness,     a.  First,  shut  the  main  valve  between 
the  pump  and  the  fire  system  lest  a  sprinkler  head  be  burst,  then  shut  all 
water  outlets  nearly,  but  not  quite  tight,  so  pump  will  move  very  slowly. 
Screw  safety  valve  down  hard.     Slowly  and  carefully  admit  steam  pressure 
sufficient  to  give  240  pounds  per  square  inch  water  pressure. 

•  b.  With  this  extreme  pressure  all  joints  should  remain  substantially  tight, 
and  the  slow  motion  of  the  pump  should  be  tolerably  smooth  and  uniform. 
(The  leakage  of  a  few  drops  here  and  there  and  a  little  unsteadiness  of  motion 
are  to  be  expected.) 

c.  If  boiler  pressure  is  above  85  pounds,  the  safety  valve  on  pump  should 
be  attached  and  screwed  down  only  enough  to  hold  the  required  pressure. 
For  with  100  pounds  or  more  of  steam  the  water  pressure  might  be  carried 
too  high. 

After  completing  the  above  test  slack  off  on  safety  valve,  setting  it  so  that 
it  will  begin  to  open  at  about  100  pounds  pressure. 

51.  Test  of  Capacity  of  Safety  Valve,    a.  The  relief  valve  may  next  be 
tested  by  first  adjusting  it  to  pop  at  100  pounds,  then  shut  the  main  outlet 
to  pump,  and  then  shut  the  hose  gates  one  by  one,  and  thus  force  all  the  dis- 
charge through  the  relief  valve,  meanwhile  opening  steam  throttle,  so  as  to 


SPECIAL  PUMPING   MACHINERY  .  495 

run  pump  first  at  two-thirds  speed  or  about  fifty  revolutions  per  minute,  and 
finally  at  full  speed  (seventy  revolutions).  The  safety  valve  (relief  valve) 
should  carry  all  this  and  not  let  the  pressure  rise  above  125  pounds. 

The  pressure  in  a  quick-moving  fire-pump  necessarily  fluctuates  5 
to  15  pounds  at  different  points  in  stroke,  and  an  air  chamber  of  reasonable 
size  cannot  wholly  remove  this.  Therefore  the  safety  valve  must  be 
set  at  about  1 5  pounds  higher  than  the  intended  average  working  pressure; 
otherwise  it  will  get  to  jumping  with  almost  every  stroke. 

52.  Test  of  Internal  Leakage  or  Slip.     a.  Set  safety  valves  at  115  pounds, 
shut  all  water  outlets,  admit  steam  enough  to  give  100  pounds  water  pressure, 
then  pump  will  move  very  slowly  under  the  influence  of  the  leakage  past 
plungers;  about  one  revolution  of  pump  per  minute  shows  a  proper  accuracy 
of  fit.    Anywhere  from  one-third  to  two  revolutions  per  minute  is  satisfactory. 

Too  tight  a  fit  is  bad,  as  if  not  exceedingly  uniform  it  induces  scoring 
or  fretting  of  the  metals.  Moreover,  should  pump  happen  to  be  run 
dry  for  a  few  minutes  before  catching  its  suction  a  slight  warming  and 
expansion  of  the  plunger  may  cause  it  to  stick  and  fret. 

53.  Test   With   Maximum   Working   Pressure,     a.  For   this,    alternately 
shut  down  the  main  outlet  gate  and  adjust  the  hand-wheel  of  the  safety  valve, 
and  open  up  on  the  throttle  as  may  be  required,  running  pump  at  say  one-half 
speed  (or,  in  experienced  hands,  at  full  rated  speed),  and  note  the  greatest 
water  pressure  which  the  full  boiler  pressure  (unless  boiler  pressure  is  above 
85  pounds)  will  yield  with  pump  at  full  speed. 

Sometimes  it  may  be  necessary  to  force  water  through  very  long 
lines  of  hose,  or  to  an  unusual  height. 

Steam  fire  engines  are  not  infrequently  called  on  to  give  200  pounds 
per  square  inch  water  pressure. 

To  test  short  hose  lines  with  anywhere  near  so  high  a  pump-pressure 
is  dangerous,  lest  nozzle  kick  and  pull  itself  away  from  the  man  holding 
it  and  thresh  around;  but  the  ability  of  the  pump  may  be  tested  by 
puttjng  this  high-pressure  delivery  mainly  through  the  safety  valve, 
or  in  part  through  the  partially  closed  main  outlet  gate. 

It  is  not  advisable  to  carry  this  water  pressure  above  200  pounds  in 
this  test  at  the  factory,  although  in  the  shop  test  the  water  pressure  is 
carried  to  240  pounds,  and  engine  driver  should  stand  with  his  hand  on 
the  throttle. 

'54.  Test  for  Maximum  Delivery,  a.  This  can  best  be  tried  by  adding 
one  or,  in  some  cases,  two  more  streams  than  the  pump  is  rated  to  deliver 
by  attaching  the  extra  lines  of  hose  to  some  hydrant  near,  and  then  speed 
up  the  pump  gradually,  to  see  how  fast  it  may  be  run  before  violent  pounding 
or  slamming  of  valves  begins. 

•  Sometimes  the  increased  delivery  can  be  drawn  off  through  an  open 
hydrant-butt,  meanwhile  holding  sufficient  back  pressure  to  show  100 
pounds  on  the  water  gauge  by  partly  closing  the  discharge  valve. 

The  engine  driver  should  stand  with  his  hand  on  or  neai  the  throttle 
when  thus  speeding  the  pump. 

It  is  all  right  to  run  a  fire-pump  up  to  the  utmost  speed  possible 
before  water  hammer  begins,  and  very  often  a  pump,  while  new  and  if 
favorably  set  up,  can  deliver  25  to  50  per  cent  more  than  rated  capacity: 
nevertheless,  although  expert  treatment  can  force  1000  gallons  from  a 
16x9x12  pump  we  can  rate  it  as  only  a  75o-gallon  pump.  There  'must 
be  some  margin  to  allow  for  wear  and  for  the  possible  absence  of  the  expert 
at  time  of  fire. 


496  PUMPING   MACHINERY 

The  main  points  of  difference  between  the  "National  Standard"  and  the 
"Trade  Pump"  are: 

Brass  plungers  instead  of  cast-iron  plungers. 

Wrought-iron  side  levers  instead  of  cast-iron; 

Bronze  piston  rods  and  valve  rods  instead  of  iron  or  steel. 

Pump  has  brass  lined  stuffing  boxes  instead  of  cast-iron. 

Rock  shafts  are  brass  bushed. 

Area  of  water  valves  is  25  to  50  per  cent  greater. 

Steam  and  exhaust  passages  20  to  50  per  cent  greater. 

Suction  pipe  connections  two  or  four  inches  greater  diameter. 

Cushion  valves  better  arranged. 

Air  chamber  is  made  much  larger. 

Shells  and  bolting  are  warranted  especially  strong. 

The  following  necessary  fittings  are  included  in  the  price,  and  regularly 
furnished  as  a  part  of  this  pump,  viz: 

A  capacity  plate. 

A  stroke  gauge. 

A  vacuum  chamber. 

Two  best  quality  pressure  gauges. 

A  water  relief  valve  of  large  capacity. 

A  cast-iron  relief  valve  discharge  cone. 

A  set  of  brass  priming  pipes  and  special  priming  valves. 

From  two  to  six  hose  valves. 

A  sight  feed  cylinder  lubricator  connected  above  throttle. 

A  one-pint  hand  oil  pump  connected  below  throttle. 

The  hydraulic  pressure  pumps  are  built  for  very  high  pres- 
sure. In  these  pumps,  as  shown  in  Fig.  387,  the  large  steam 
cylinders  are  connected  with  small  water  cylinders  GG.  .  In 
the  pump  shown  steam  is  used  in  the  cylinder  E  and.  then 
discharged  into  cylinder  F.  The  piston  rods  A  A  are  connected 
to  the  cross  head  B  and  from  this  the  plungers  extend  into  the 
two  water  cylinders.  The  connecting  rod  C  is  forked  at  DD 
so  that  it  clears  the  right  hand  pump  G.  The  use  of  the  fly- 
wheel permits  of  the  expansive  use  of  steam.  Such  pumps 
are  often  connected  to  a  cylinder  closed  by  a  weighted  ram 
(an  accumulator),  so  that  the  water  may  be  stored  under  pres- 
sure which  may  amount  to  several  thousand  pounds  per  square 
inch.  The  use  of  an  accumulator  permits  the  pump  to  run 
continuously  even  though  the  use  of  water  is  intermittent. 
When  there  is  such  a  load  that  the  pump  will  have  a  number 
of  periods  of  inaction,  a  direct-acting  pump  is  better  than  a 
fly-wheel  pump  with  a  fixed  stroke,  as  there  is  then  little  or  no 


FIG.  387.— Pres 


3ump. 


(To  face  page  496} 


• 


SPECIAL  PUMPING  MACHINERY 


497 


danger  of  the  pump  being  wrecked  from  a  collection  of  con- 
densation in  the  stearn  cylinder.  The  valve  boxes  HH  are 
similar  to  the  pressure  valve  boxes  shown  earlier. 

At  times  triplex  pumps  are  used  for  this  purpose.  Such 
pumps,  Fig.  388,  are  made  of  heavy  parts.  The  water  is  pumped 
directly  into  the  ram  cylinder  and  from  this  it  is  discharged  to 
the  reservoir.  A  small  by-pass  valve  is  set  to  discharge  from 


FIG.  388. — 5oo-ton  Press  and  Pump. 

the  pump  to  the  reservoir,  when  the  hand  valve  at  the  press 
is  shut  off  or  as  soon  as  the  pressure  in  the  system  reaches  a 
certain  amount.  The  15  horse- power  motor  used  with  this 
pump  of  the  Waterbury  Farrel  Foundry  and  Machine  Co.  is 
directly  connected  to  the  pump,  although  at  times  belted  con- 
nections are  used.  The  proportions  of  the  pump  and  press 
are  given  in  the  table  below,  which  is  taken  from  the  catalogue 
of  the  builders. 


498 


PUMPING  MACHINERY 


Press. 

Pump. 

Capacity  to 

ns       500 
i.            16 

5 

22 

40 

18^  by  12^ 
26 

165  by  .36 

82i 
S.      I7I7S 

Diameter  of  plunger  

...in                ij 
6 

Maximum  stroke  of  ram  
Diameter  top  of  ram 

Approximate    capacity, 
inches  
Pressure,  Ibs.  per  sq.in   . 

cubic 
1060 
4975 

Top  platen  to  ram  when  down 
Opening  between  rods  
Floor  to  top  of  ram  
Size  of  supply  pipe  
Floor  space,  with  pump  .... 
Extreme  height  from  floor  .  .  . 
Total  weight,  including  motor.  It 

R.P.M.  of  crank-shaft.  .  . 
Rafio  of  gearing 

61* 

Floor  space  

.  .  .  in.    91  by  32 

Figs.  389  and  390  show  two  air  pumps  for  beer  racking. 
These  pumps  are  mounted  on  the  air  reservoir  which  acts  as  a 


FIG.  389. — Worlhington  Beer-racking  Pump.     (Size  5^X6X5.) 

bed  plate.  This  gives  proper  storage  capacity.  The  cylinders 
are  lined  with  a  composition.  The  regulator  shown  on  each 
pump  is  connected  with  the  air  reservoir  and  keeps  the  pressure 
constant.  The  pump  of  Fig.  390  shows  how  a  fly  wheel  can  be 
used  between  the  cylinders  of  the  steam  and  air  ends.  A  yoke 
is  placed  in  the  rod  between  the  two  cylinders  and  a  crank  pin 
of  the  shaft  operates  between  these.  This  is  the  equivalent 


SPECIAL  PUMPING  MACHINERY 


499 


of  a  connecting  rod  of  infinite  length,  and  by  its  use  the  fly  wheel 
may  be  applied  so  that  steam  can  be  employed  expansively. 


FIG.  390. — Beer-racking  Air  Pump. 

The  table  below  gives  the  sizes  of  this  type  of  pump  as  made  by 


Worthington : 


1 

* 

0) 

1 

Sizes  of  Pipes  for 
Short  Lengths 
to  be  Increased  as 
Length  Increases. 

Approximate 
Space  Occupied. 
Feet  and  inches. 

*o  52' 

"o  £ 

t/2 

Annticil  C/3,pa,city  of 
Brewery. 

s! 

1.1 

.£3 

<u 

j§  fl) 

g« 

0  « 

jA 

S^ 

B7, 

1 

IK 

|£ 

.s  a 

^^ 

M 

D 

1 

Q 

Q 

j 

w 

Q 

JS 

'P 

5l 

si 

5 

20,  ooo  to    5o,ooobbls. 

f 

«i 

!l 

! 

3      7 

1      9 

6 

8 

6 

50,  ooo  to  ioo,ooobbls. 

i 

ii 

2 

1^ 

4      9 

2      4 

7i 

9 

10 

100,  ooo  to  i5o,ooobbls. 

Ji 

2 

3 

3 

6      7 

2      5 

9 

12 

10 

1  50,000  to  200,000  bbls. 

2 

*\ 

4 

3 

6    10 

3      o 

To   designate   the   sizes,    give   the   diameters   of   the   steam   and   air 
cylinders  and  the  length  of  stroke. 


500 


PUMPING   MACHINERY 


The  hydraulic  ram  in  one  of  its  modern  forms  is  shown  in 
Fig.  391,  while  Fig.  392  illustrates  the  method  of  connecting 
it.  The  drive  pipe  F  connects  the  source  of  supply  A  with  the 
ram  at  B  at  a  lower  level.  The  water  can  escape  from  the  ram 
by  an  opening  or  waste  valve  K.  The  velocity  set  up  at  the 
discharge  opening  is  sufficient  to  lift  the  valve  G  and  suddenly 


FIG.  391. — Hydraulic  Ram. 
\ 

close  the  passage.  The  momentum  of  the  water  in  the  pipe  line 
produces  an  increase  in  the  pressure  near  the  valve,  as  shown 
from  the  rise  from  a  to  b  in  Fig.  393.  This  figure  was  drawn 
by  the  pencil  of  an  indicator  attached  to  the  pipe  F,  known  as 
the  drive  pipe,  while  the  drum  was  moved  uniformly.  The 
heights,  therefore,  represent  pressure  while  length  represents 
time.  The  increase  of  pressure  will  cease  as  soon  as  the  pressure 
is  sufficiently  great  to  open  the  ball  valve  H  when  the  water 


SPECIAL  PUMPING   MACHINERY 


501 


flows  into  C.  The  inertia  and  friction  of  the  valve  being  over- 
come, the  pressure  drops  with  a  backward  wave  to  c  and  returns 
to  d,  after  which  the  ball  valve  closes;  the  pressure  drops  in  the 
space  around  G  and  a  backward  wave  may  even  reduce  this  to  a 
pressure  below  the  atmosphere.  In  any  case  the  weight  /  is 


FIG.  392. — Arrangement  of  Ram. 

sufficient  to  force  G  down  against  the  water  pressure  due  to 
the  elevation  of  A,  after  which  the  operation  is  repeated  at 
g,  h,  i,  etc.  The  discharge  is  taken  from  L  by  the  discharge 
pipe  D  which  conducts  it  to  the  tank  house. 

In  making  and  installing  the  ram,  the  design  of  pipe  and 
size  is  fixed  by  practise,  although  there  are  theoretical  deductions 
made  at  times.  (Zeitschrift  des  Vereines  Deutscher  Ingenieure, 


h  9  e     a- 

FIG.  393. — Card  from  Hydraulic  Ram. 

Jan.  15,  1910.)  The  drive  pipe  F  should  be  of  ample  size 
and  should  be  laid  as  straight  as  possible.  It  should  be  of  a 
length  at  least  equal  to  three-quarters  the  head  against  which  it 
is  to  pump,  or  five  times  the  head  causing  the  flow  through  the 
drive  pipe.  The  inclination  of  the  drive  pipe  should  never  be 
over  30°  at  any  place,  and  when  the  length  or  proper  angle 
cannot  be  had  it  is  well  to  coil  the  drive  pipe  to  a  large  radius  so 


502  PUMPING  MACHINERY 

as  to  accomplish  this.    The  length  of  the  pipe  is  necessary  to 
produce  the  momentum  for  driving.     The  length  of  delivery 
pipe  should  not  be  over  twenty  times  the  lift  and  when  a  longer 
pipe  must  be  used  less  lift  will  be  obtained. 
The  efficiency  of  the  hydraulic  ram  is 

w(H-h) 
~Wh~' 
where 

#=head  from  ram  to  reservoir  in  feet; 
/j=head  from  supply  to  ram  in  feet; 
/  =  time  in  seconds; 
w—  water  pumped  in  Ibs; 
W  =  water  discharged  through  waste  valve  in  Ibs. 

This  efficiency  varies  from  o .  40  to  o .  60,  and  by  assuming  o .  50 
for  its  value,  the  equation  may  be  used  to  find  W  when  w,  H 
and  h  are  known.  The  drive  pipe  should  then  be  made  large 
enough  to  give  a  velocit}^  of  three  feet  per  second  and  the 
delivery  pipe  two  feet  per  second. 


62.5X/X3  ' 
^del-=62.5X/X2' 


CHAPTER  XII 

. 

INJECTOR   AND   PULSOMETER 

THE  Sellers  injector,  Fig.  394,  has  been  taken  as  a  type  of 
the  injector.  In  starting,  the  handle  A  is  drawn  back  a  short 
distance  permitting  steam  to  enter  the  space  B  between  the 
two  tubes.  The  pressure  of  this  steam  between  the  two  edges 
produces  so  high  a  velocity  of  discharge  across  the  upper  end 
of  space  C  that  it  entrains  the  air  at  that  point,  driving  it  out 
through  the  openings  D  E  into  the  space  G,  and  from  this 


FIG.  394. — Sellers  Injector. 

into  the  atmosphere.  This  produces  a  partial  vacuum  at  C 
and  water  enters  through  the  suction  pipe  H.  This  water  is 
taken  up  by  the  steam  discharge,  driven  through  the  combining 
tube  /  and  is  afterwards  discharged  through  Z),  and  E,  finally 
appearing  at  the  overflow. 

When  water  appears  at  the  overflow,  the  handle  A  is  drawn 
back  completely  and  steam  then  discharges  through  the  main 
nozzle  K.  The  steam  in  this  acquires  a  high  velocity  due  to 

503 


504  PUMPING  MACHINERY 

the  drop  in  pressure  in  the  nozzle;  this  velocity  is  many  times 
greater  than  'that  acquired  by  water  under  the  same  drop  in 
pressure  on  account  of  the  small  density  of  the  steam. 

The  steam  is  then  condensed  by  the  water,  but  when  the 
steam  particles  draw  together  to  form  a  drop  of  water  their 
velocity  is  maintained  and  drops  of  water,  moving  with  a 
very  high  velocity,  result.  This  condensed  steam  can  then 
strike  the  body  of  water  drawn  through  H  and  by  impact, 
impart  to  it  a  velocity  greater  than  that  which  would  be  acquired 
by  water  discharging  from  the  same  boiler.  If  this  is  the  case, 
this  mixture  of  water  and  condensed  steam  could  enter  the 
boiler  against  steam  pressure. 

The  combining  tube  7  is  made  convergent  since  the  steam 
is  gradually  condensed  in  the  passage  and  the  water  gradually 
increases  in  velocity.  When  the  throat  L  is  reached  the  steam 
is  condensed  and  the  water  is  supposed  to  have  its  maximum 
velocity.  From  this  point  the  delivery  tube  M  diverges,  making 
the  velocity  less,  and  with  this  comes  a  change  of  velocity  head 
into  pressure  head  until  at  the  end  there  is  sufficient  pressure 
to  force  open  the  valve  0  when  the  mixture  will  enter  the  boiler. 

The  pressure  in  the  space  within  the  end  of  the  combining 
tube  reaches  a  point  of  low  vacuum  of  about  25  "  while  at 
the  end  of  the  steam  nozzle  it  is  at  atmospheric  pressure.  At 
the  small  part  of  the  nozzle  known  also  as  the  throat,  the  pres- 
sure is  found  to  be  0.58  of  the  boiler  pressure.  The  pressure 
at  the  throat  of  the  delivery  tube  is  usually  about  atmospheric 
pressure. 

If  how  the  pressure  of  the  steam  is  p  pounds  per  square 
inch,  absolute,  and  the  barometer  reads  pa,  the  pressure  against 
which  the  water  must  enter  is  (p  —  pa)  and  if  ^  is  the  density  of 
the  water  or  weight  per  cubic  foot  at  the  throat,  and  v-2  the  veloc- 
ity entering  the  boiler,  the  head  in  feet  of  water  to  be  overcome  is 


2g 


A  may  be   taken   as   50   pounds   for  the   first  approximation. 
JTie  reason  for  this  low  density  is  the  fact  that  steam  particles 


INJECTOR  AND  PULSOMETER  505 

are  mixed  with  the  water.     Such  a  result  has  been  found  from 
experiment.     The  velocity  at  the  point  E  will  be 


Vd  = 

The  velocity  of  water  drawn  through  the  suction  tube  is 
produced  by  the  pressure  pc  within  the  combining  tube,  and 
if  the  water  is  lifted  h  feet  and  the  friction  is  h/  feet  the.  head 
causing  flow  is 


62.5 

the  velocit    of  the  water  is 


The  steam  discharging  from  the  nozzle  is  under  a  pressure 
of  p  originally,  and  in  the  combining  tube  it  is  under  a  pressure 
of  pe.  To  find  the  velocity,  the  heat  content  H  of  the  initial 
steam  is  found,  considering  the  quality  of  the  steam  and  then 
the  heat  content  Hc  at  the  lower  pressure,  assuming  that  the 
steam  has  the  same  entropy  at  the  two  points.  If  15  per  cent 
friction  loss  is  assumed,  the  velocity  V  of  the  steam  is  given 
by  the  equation: 

S52-£[H-Hc]    where    H  =  q+xr  and  A  =  -±-. 

Having  now  the  velocities  of  the  steam,  water  and  mixture, 
the  equation  for  impact  may  be  used  to  find  the  amount  of 
water  per  pound  of  steam. 


Since  impact  is  not  perfect,  the  sum  of  the  first  products  has  to 
be  multiplied  by  a  factor  K.  Mr.  Strickland  L.  Kneass,  in  his 
"  Theory  of  the  Injector,"  has  shown  that  a  value  of  0.50 
should  be  used  for  K  in  the  formula: 


KV-Vd 


506  PUMPING  MACHINERY 

With  M  known,  the  condition  of  the  mixture  can  be  found  by 
equating  the  heat  on  each  side  of  the  throat  in  which  qd  is  the 
onl  unknown. 


of  liquid  in  suction, 
2^=  heat  of  liquid  of  discharge, 
H  =  total  heat  content  of  entering  steam. 

The  density  of  the  water  in  the  discharge  depends  on  the 
temperature  of  the  water  and  the  amount  of  steam  left  in  the 
water  at  this  point.  Kneass  has  shown  how  this  may  be  found, 
and  in  some  cases  its  value  is  o.2iw,  where  w  is  the  weight  of 
a  cubic  foot  of  water  at  the  temperature  of  discharge.  This 
result  was  determined  for  a  small  discharge.  The  value  0.7520 
is  more  nearly  the  value  to  be  used  in  practice.  This  would 
give  45  pounds  with  water  weighing  60  pounds  per  cubic  foot. 

If  the  weight  of  water  wanted  per  hour  is  W  pounds,  the 
amount  of  steam  per  second  is 


6ox6oM 

hence  the  orifice  must  be  of  area  sufficient  to  cany  w'(i+M) 
pounds  of  density  A  at  the  speed  of  Vd  feet  per  second.  The 
area  at  the  throat  is  therefore 


The  diameter  of  this  expressed  in  millimeters  gives  the 
nominal  size  of  the  injectors  according  to  some  makers. 

The  area  of  the  throat  of  the  steam  nozzle  is  sometimes 
made  two  or  three  times  the  area  of  the  delivery  throat,  but 
this  may  be  designed  by  considering  the  velocity  set  up  by 
the  fall  of  pressure  from  p  to  0.57/7.  Let  H  be  the  heat  content 
for  the  steam  under  the  conditions  at  entrance  and  Ht  that 
at  0.57/7,  but  with  the  same  entropy,  then 


778' 


INJECTOR  AND  PULSOMETER  50? 

The  factor  0.98  is  used  here  because  of  the  short  easy  curve. 


A  -1 
Vt 

St  is  the  specific  volume  of  the  steam  at  the  throat. 

If  He  is  the  heat  content  of  the  steam  at  the  end  of  the 
nozzle  where  the  pressure  has  fallen  to  about  atmospheric 
pressure,  the  following  may  be  used  to  find  the  area  at  the  end: 


V, 

He  is  taken  from  the  same  entropy  column  as  H  but  in  finding 
Ve,  the  volume  is  that  which  corresponds  to  a  heat  content 
H'e=He+Q.i5(H-He).  The  reason  for  this  is  the  fact  that 
the  loss  due  to  friction  has  been  changed  into  heat  and  remains 
in  the  steam  increasing  the  heat  content  over  that  resulting 
from  true  adiabatic  expansion. 

The  nozzle  is  usually  rounded  to  the  throat  area  with  quad- 
rants and  then  the  remainder  of  the  nozzle  is  made  along  a 
straight  line  to  the  end.  The  combining  tube  slopes  from  the 
delivery  throat  to  an  area  somewhat  larger  than  the  nozzle  end. 

The  delivery  tube  should  be  made  of  gradual  convergence, 
and  Kneass  recommends  ("Theory  of  Injectors,"  p.  37)  to  use 
a  curve  such  that  the  negative  acceleration  along  the  tube 
is  constant.  This  gives  the  equation: 


4   /" 


v  V2-2ay 

Where  R  is  the  radius  of  the  tube  at  a  distance  y  from  the  throat 
of  radius  r  and  a  is  the  acceleration  while  V  is  the  velocity  at 
the  throat. 

Suppose  that  the  velocity  is  to  be  destroyed  in  /  feet  of 
tube,  then 

F2 


508 


PUMPING  MACHltf&tV 


This  may  be  put  in  the  above  equation  and  the  following 
results: 


from  this  the  various  radii  may  be  found  along  the  tube. 

The  shapes  of  the  tubes  actually  used  in  practice  have  been 
determined  by  experiment,  and  the  above  serves  as  a  guide  for 
certain  leading  dimensions. 

The  pulsometer  described  in  Chapter  II  is  built  under  a 
number  of  names  and  with  various  shapes,  but  of  the  same 
principle.  Fig.  395  shows  the  Emerson  Pump.  The  cylindrical 
vessels  A  A  are  for  the  same  purpose  as  the  pear-shaped  vessels 
of  the  pulsometer. 

The  pulsometer  is  operated  by  the  action  of  the  steam  on 
the  water  direct.  The  quantity  of  water  to  be  handled  by  a 
given  size  is  fixed  from  practice  and  the  table  below  gives  these 
sizes  taken  from  a  catalogue. 

TABLE  OF  SIZES  OF  EMERSON  STEAM  PUMPS 


e 

*i2 

Dimensions  Over 

2 

^ 

T3 
C 

1 

i 

£ 

§" 

8, 

§ 

"d 

1 
'rj 

II 

all  in  Inches. 

1 

O 

">. 

B 

. 

rt 

,c 

Ccj 

0 

O  J 

0 

•a 

o 

i 

ts 

jj 

C3 

c  jj 

S  -=: 

^-> 
rt 

Number. 

Diameter 
Inches. 

Length  of 
Feet. 

i  SizeofStc 
Inches. 

|l 

u 

ft 

P 

1! 
|l 

y 

Js 

o 

***    rS 

&3 

u 

O 

Breadth. 

Width. 

^3 
I 

Approxim 
in  Ibs. 

I 

6 

6 

a 

4 

3 

•1 

225 

i3.5oo 

324,000 

i6i 

18 

97* 

95° 

2 

8 

6i 

I 

4 

3 

415 

24,900 

597,600 

21^ 

21 

104 

1370 

3 

10 

7 

II 

•  5 

4 

725 

43.50° 

1,044,000 

26 

24 

"3 

J905 

4 

12 

8 

I* 

6 

5 

1200 

72,000 

1,728,000 

29! 

27i 

127 

3100 

5 

16 

8 

2 

8 

6 

2IOO 

126,000 

3,029,000 

43  i 

33 

132 

4400 

6 

20 

8 

»l 

10 

8 

3275 

196,500 

4,7l6,OOO 

5*1 

36J 

!35 

5400 

7 

24 

8 

3 

12 

10 

4700 

282,000 

6,768,000 

7000 

Capacities  in  gallons  per  minute,  stated  in  table,  vary  with  the 
steam  pressure  and  height  of  lift.  Special  sizes  above  those  listed, 
on  application.  Pumps  made  entirely  of  bronze,  when  called  for. 
at  special  prices. 


INJECTOR  AND  PULSOMETER 


509 


IP 

r 


FIG,  395. — Section  of  the  Emerson  Steam  Pump. 


510  PUMPING  MACHINERY 

To  compute  the  quantity  of  steam  used,  suppose  that  steam 
is  supplied  at  a  pressure  pi  and  is  used  within  the  chamber 
at  pressure  p2.  Assume  that  there  is  no  radiation  and  if  this 
is  the  case  the  steam  in  passing  through  the  valves  is  reduced 
in-  pressure  but  the  heat  content  remains  the  same.  Hence 

HI  =H2. 

From  HZ  at  the  pressure  of  p2,  the  specific  volume  S2  of  the 
steam  can  be  obtained  and  then  for  a  given  quantity  of  water, 
V  cu.ft,  the  weight  of  steam  used  would  be 

V 

—  —  =  wi. 

Oo 

'      ,  -. 

The  heat  in  each  pound  of  this  steam  is  H  =  q±+Xiri,  and  if 
the  temperature  of  discharge  is  td  the  heat  chargeable  per  pound 
of  steam  is  H—qd. 

qd  and  the  temperature  of  the  mixture  is  given  by  the  equation 


Mqa  +m(qi  +x^i  )  =  (M  +  m)qd  +  A  144^2  (M  + 

V 

M  =weieht  of  V  cu.ft.  of  water  =  7  —  , 

62.5' 

of  liquid  in  suction. 


=  volume  of  I  Ib.  of  w;iter=z  —  . 

62.5 


The  efficiency  of  the  pump  is 
work 


heat      7?8m(H-qdY 

The  above  equations  have  considered  no  loss  due  to  initial 
condensation  nor  radiation,  and  these  depend  on  many  con- 
ditions. It  may  be  said  that  the  steam  will  probably  be  increased 
to  $m  or  more  and  this  will  reduce  the  efficiency  to  about  one 


INJECTOR  AND  PULSOMETER  511 

third  its  former  value.     The  term  q^  will  be  changed  as   its 
equation  becomes 

Mqs  +  3m(qi  +xtfi )  -=  (M  +^m)q  -{-Radiation  +Work. 
Radiation  =K(A  )(T±  - T2 )t\ 

K  =  300  =  amount  of  heat  radiated  per  sq.ft.  per  hour 

per  degree  difference  in  temperature; 
A  —area  of  outside  of  vessels; 
TI  --=  temperature  on  inside  of  pulsometer; 

__  J-  steam  "r  J-  d 
2 

T2  =  temperature  air; 
^  =  time  in  hours  in  which  V  cu.ft.  of  water  is  pumped. 


CHAPTER  XITI 
AIR   LIFT   PUMPS  AND  PNEUMATIC    PUMPS 

THE  air  lift  pump,  as  was  mentioned  in  Chapter  II,  was 
patented  in  1880  by  James  P.  Frizell,  and  Pohle  took  out  a 


FIG.  396. — Air  Lift. 

patent  in  1886,  although  these  two  are  by  no  means  the  first 
records  of  this  type  of  pump. 

Fig.  396  .shows  the  general  arrangement  of  a  plant  for  an 


512 


AIR  LIFT  PUMPS   AND   PNEUMATIC  PUMPS  513 

air  lift.  In  this  figure  the  air  pump  is  attached  to  a  storage 
tank  from  which  the  air  is  conducted  to  the  well.  The  air 
compressor  is  shown  to  be  of  a  single  stage  type  although  in 
many  cases  two  stage  compressors  are  used  to  increase  the 
efficiency.  The  air  cylinder  A  is  in  tandem  with  the  steam 
cylinder  B.  The  air  from  the  storage  tank  C,  with  its  gauge  and 
safety  valve,  is  conducted  to  the  head  of  the  discharge  pipe 
at  E  through  the  pipe  D.  This  pipe  is  continued  down  to  a  point 
F  near  the  bottom  of  the  delivery  pipe.  The  well  is  usually 
cased  outside  of  the  delivery  pipe  through  all  earthy  matter 
to  solid  rock  if  certain  ground  waters  are  to  be  kept  from  the 
well.  The  water  in  the  well  stands  at  a  height  h  below  the 
discharge  and  the  depth  of  immersion  from  G  to  H  is  called  h'. 

The  figure  shows  several  methods  of  introducing  the  air. 
In  the  first  place  the  pipe  is  carried  down  inside  of  the  main 
pipe  and  opens  'directly  into  the  delivery  pipe.  There  are  two 
other  figures  showing  the  small  pipes  carried  down  into  the 
well  and  introduced  into  the  side  of  the  delivery  pipe,  while  the 
fourth  figure  shows  the  method  of  carrying  the  air  down  in 
an  annular  space  between  two  pipes,  the  inner  one  of  which 
is  the  delivery  pipe.  A  slide  valve,  controlled  from  the  top 
of  the  well,  admits  air  at  some  height  above  the  bottom  for  the 
purpose  of  introducing  air  at  a  higher  point  when  starting  the 
apparatus.  The  air  in  this  case  is  introduced  at  I  into  a  head 
at  the  well  top. 

There  are  several  methods  of  arranging  the  well  tops  for 
the  reception  of  the  discharge  and  for  the  introduction  of  the 
air  pipes.  These  are  shown  in  Fig.  397.  In  the  first  a  concrete 
head  A  receives  the  discharge  from  the  deflecting  cap  G.  It 
is  carried  away  from  this  by  the  conduit  leading  to  the  reservoir 
or  irrigation  ditch.  The  handle  at  H  controls  the  admission 
of  the  air  by  a  rod  which  extends  to  the  lower  end  of  the  air 
pipe.  When  the  water  is  to  be  carried  to  a  higher  level,  the 
S  bend  shown  at  B  is  used,  but  the  control  valve  entering  the 
pipe  at  the  ground  level  is  similar  to  that  used  in  A.  When 
an  elbow  is  used,  the  connection  for  the  air  pipe  is  shown 
at  C  and  the  forms  at  D  and  E  are  those  used  with  the  "i* 


514 


PUMPING   MACHINERY 


supplied  in  the  annular  space  outside  of  the  delivery  pipe. 
In  E  the  discharge  is  caught  in  a  tank  before  being  delivered, 
while  in  D  the  discharge  is  directed  into  the  irrigation  ditch 
or  cistern. 

Such  installations  may  be  arranged  at  considerable  distance 
from  a  central  air  compressing  station.  One  station  is  used 
to  furnish  air  to  a  number  of  wells.  The  pipe  lines  should 
be  arranged  so  that  the  air  travels  at  a  rate  of  from  2000  to 
4000  feet  per  minute.  The  area  of  this  pipe  is  usually  cne- 
sixth  the  area  of  the  delivery  pipe.  Some  pump  makers  claim 
that  this  method  of  raising  water  should  not  be  applied  when 
the  water  has  to  be  raised  over  80  feet,  but  others  name  180  to  200 
feet  as  the  limit,  and  when  the  water  is  to  be  carried  by  the  air 


FIG.  397.— Air  Lift  Well  Tops. 

pressure  in  addition  to  a  small  lift  the  distance  is  limited  to 
700  or  800  feet.  Be  this  as  it  may,  greater  lifts  are  used,  although 
the  cost  of  lifting  may  be  excessive. 

When  greater  lifts  are  required  it  might  be  well  to  use 
multiple  lifts  discharging  through  a  portion  of  the  height  to 
a  deep  pipe  reservoir  in  the  main  well  and  from  this  to 
another. 

The  compressors  should  in  all  cases  be  arranged  to  compress 


AIR  LIFT  PUMPS  AND   PNEUMATIC  PUMPS 


515 


the  air  with  as  nearly  isothermal  compression  as  possible.  The 
only  way  to  do  this  is  to  use  an  inter-cooler  between  the  stages. 
When  the  pressure  is  from  60  to  300  pounds,  two  stages 
should  be  used,  and  above  this  three  or  more  stages.  The  air 
pipe  should  be  introduced  near  the  bottom  of  the  discharge 
pipe  and  should  be  immersed  so  far  that  the  ratio  of  h' 
to  h  is  3  to  i  at  the  start  and  2.2  to  i  in  operation,  according 
to  a  test  on  a  particular  well  to  find  out  what  depth  of  im- 
mersion gave  the  best  results.  This  result,  however,  should 
not  be  used  as  a  general  law  as  there  are  many  conditions 
affecting  it.  Some  claim  that  the  ratio  should  be  1.2:1  to  4:  i. 
These  will  do  as  guides.  The  following  table  gives  a  series  of 
tests  showing  what  may  be  expected: 


Place. 

h' 

Immersion  -=- 
ft 

Efficiencies. 

Tunbridge,  England  

3:1  to  2.2:1 

?6% 

Grinnell,  Iowa 

2Q   6% 

San  Francisco  

0.6  :  i 

16%     4-1% 

San  Francisco 

i  '  i 

IO%        42% 

San  Francisco  

1.4  :  i 

74%        41% 

San  Francicso 

2   A.  '  I 

T  r%        24% 

San  Francisco 

•}     Q  '   I 

2% 

j    e 

CO 

o" 

o  66  to  i 

o  ^o 

JV 

2  C. 

0.43  to  i 

The  efficiency  shown  in  this  table  is  the  ratio  of  the  work 
done  on  the  water  to  that  done  in  compressing  the  air. 

The  pressure  to  be  carried  on  the  air  system  is  greater  than 
the  water  pressure  at  the  lower  end  of  the  discharge  pipe  or 
p  =0.434/^+5.  This  quantity  diminishes  after  the  pump  starts 
to  act  as  the  head  h'  decreases,  due  to  the  removal  of  water. 
The  difference  between  the  head  at  start  and  while  running, 
represents  the  head  causing  flow  into  the  well.  This  quantity 
will  vary  in  different  localities,  depending  entirely  on  the  form- 
ation and  the  nature  of  the  water  bearing  rock. 

The  amount  of  free  air  varies  according  to  different  authors. 


516 


PUMPING  MACHINERY 


One  gives  the  amount  as  3.9  cubic  feet  to  4.2  cubic  feet  of 
free  air  per  cubic  foot  of  water,  or 

cu.ft.  of  free  air  per  min.  =   f        . 

16.824 

L  =  lift  of  water  above  surface  in  well  in  feet. 
Q  =  cu.ft.  of  water  per  min. 

A  pump  manufacturer  recommends  the  following: 

cu.ft.  air  =  — , 
while  another  uses: 

—  =  cu.ft.  of  air  per  min. 

The  air  is  transmitted  at  2000  feet  per  minute,  and  the 
c~rea  of  water  pipe  should  be  about  six  times  that  of  the  air 
pipe.  The  immersion  should  be  i±  times  the  lift.  The  table 
below  has  been  given  in  "The  Engineer"  for  June  i,  1906: 


Size  of  Well. 

W*ater  Pipe. 

Air  Pipe. 

Gallons  per  Minute. 

4" 

ii 

I 

25 

44" 

2 

I 

5° 

5" 

24 

I 

75 

6" 

3 

it 

IOO 

7" 

3i 

J4 

150 

8" 

4 

li 

200 

9" 

5 

2 

300 

10" 

6 

2 

45° 

This  same  article  gives  the  following  relation  for  the  cubic 
feet  of  free  air  per  cubic  foot  of  water: 


Lift. 

Cu.ft.  Air  per 
Cu.ft.  Water. 

.  Submergence 
if  1.5. 

Air  Pressure  in  Ibs. 

25 

2 

37-5 

17 

5° 

2 

75 

33 

75 

IOO 

I25 
150 

4-5 
6 

7-5 
9 

112.5 

!5° 

187.5 
225 

49 
65 
82 
98 

J75 

IO-5 

202  .  5 

JI5 

200 

12 

300 

130 

AIR  LIFT  PUMPS  A21D  PNLUI.IATIC  PUMPS 


517 


As  an  example,  the  following  data  may  be  mentioned: 
A  lift  of  129  feet  was  obtained  from  a  well  300  feet  deep;  the 
water  was  44  feet  from  the  top  of  the  well,  leaving  an  85-foot 
lift  above  the  well.  The  well  was  8  inches  in  diameter  and 
the  discharge  pipe  was  3^  inches  with  a  ij-inch  air  pipe.  The 
pump  gave  82.5  gallons  per  minute  and  required  7.43  cubic 
feet  of  air  per  cubic  foot  of  water.  The  air  pressure  required 
was  107  pounds  and  the  loss  in  pressure  was  9%. 

As  another  example  of  the  use  of  the  air  lift  pump, 
the  plant  at  Redlands,  California  will  be  mentioned.  This 
was  described  in  the  "Engineering  Record,"  Vol.  51,  p.  8. 
The  plant  was  built  to  replace  machinery  at  four  wells,  from 
450  to  570  feet  deep,  which  were  separated  some  distance.  The 
first  pump  was  of  centrifugal  form  and  was  operated  by  a  motor; 
the  second,  a  plunger  pump  driven 
by  a  steam  engine;  the  third,  a 
centrifugal  pump  with  a  steam 
engine,  and  the  fourth  had  not 
been  used.  It  was  decided  to 
operate  these  wells  by  air  from  a 
central  station.  The  wells  were 
to  be  piped  with  4-inch  and  7- 
inch  discharge  pipes,  extending 
down  from  306  to  360  feet,  while 
the  air  pipes  were  ij  to  2  inches 
in  diameter. 

The  station  was  a  brick  build- 
ing 40X46  feet  and  equipped 
with  return  tubular  boilers  66 
inches  in  diameter  and  16  feet 
long.  These  used  oil  for  fuel.  A 
13  and  26x30  Cross  compound 

engine  with  14  and  22X30  two-stage  air  cylinders  in  tandem, 
supplied  1124  cubic  feet  of  free  air  per  minute  to  125  pounds 
at  85  R.P.M.;  at  this  time  it  developed  190  I.H.P.  The  boiler 
feed  pumps  at  the  station  were  a  2fx  4-inch  duplex  pump  and 
a  4!  X3X 4-inch  triplex  pump. 


FIG.  398.— Wheeler  System. 


518  PUMPING  MACHINERY 

The  48-hour  test  on  this  .plant  gave  the  following  results: 


Mean  for  48 
Hours. 

Maximum. 

Minimum. 

Boiler  pressure          

146  .  7 

I  SO 

14.2 

Air  compressor  pressure 

QO    08 

04. 

8c 

Vacuum  in  inches     

24.83 

2  C  .  2  C 

°o 

24.     £ 

Well  No.  i  : 
Air  pressure 

8?     A 

80 

•^  •  0 
SA 

Depth  water  level             

101    44 

I  OQ 

08 

Well  No.  2: 
Air  pressure                   

Q-2   .  6 

0  S 

Depth  water  level  

QO  .  77 

Q-? 

86 

'Well  No.  3: 
Air  pressure 

87  .01 

QO 

8s 

Depth  water  level            

I  IO    67 

116 

Well  No.  4: 
Air  pressure 

8c   4. 

86 

SA 

Depth  water  level           

IO2     2 

JOS 

IOI 

Fuel  consumed  in  barrels  per  24  hrs  . 
Rate  of  pumping  in  gals,  per  24  hrs  . 

I6.3 

3,157,622 

17.2 
3,280,824 

15-7 
3,i3i.395 

When  sufficient  immersion  can  not  be  had,  the  Wheeler 
system,  Fig.  398,  may  be  used.  Water  is  admitted  into 
one  of  the  two  vessels  A  by  drawing  out  the  air,  when  water 
will  flow  in  from  the  well;  then,  on  allowing  air  pressure  to 
enter  through  B,  this  forms  an  equivalent  head  on  the  water. 

The  action  of  the  pump  and  compressor  will  now  be  exam- 
ined. Suppose  that  v  cubic  feet  of  air  are  admitted  per  cubic 
foot  at  the  bottom  of  the  discharge  pipe,  under  a  pressure  of 
feet.  The  specific  gravity  of  the  mixture  of  i  cubic 


foot  of  water  and  v  cubic  feet  of  air  will  be  -  —  ,  neglecting  the 

I  Til 

weight  of  air. 

The  pressure  of  the  mixture  in  the  pump  pipe  is  h'  +34  feet 
of  water  at  the  bottom  of  the  pipe  and  34  feet  at  the  top  after 
rising  through  h+h'  feet  of  the  pipe.  This  pressure  is  equal 
to  the  weight  above  this  point  of  the  column  of  mixture  of  one 
square  foot  cross-section  plus  the  atmospheric  pressure.  It 


AIR  LIFT  PUMPS  AND  PNEUMATIC  PUMPS  519 

will  vary  according  to  some  law  which  must  be  found.  It  is 
evident,  however,  that,  if  it  be  assumed  that  the  air  is  not 
absorbed  by  the  water,  the  volume  of  the  air  will  vary  inversely 
as  the  pressure  at  any  point  considered,  because  there  is  so 
much  water  in  contact  with  the  air  that  the  expansion  of  the 
air  is  isothermal.  At  any  point  x  the  pressure  is  that  due  to 
the  weight  of  a  column  of  mixture  x  feet  high  plus  the  atmospheric 
pressure  of  34  feet  of  water.  Call  the  weight  of  the  column 
whx  if  expressed  in  feet  of  standard  water. 

,        Cx  wdx 

whx  = 


-'-, 

Jo 
Now  vx 


Differentiating  the  expression  for  whx, 

j7       wdx 
wdhx  =  — = — , 

dx 


dhx  = 


I+,^±34)' 
(A.+34) 


fe-f34 

x 
When  x  =  o,  hx  =  o',  and  when  x  =  h'+hy  hx  =  ti. 

.'.     c  =  o-v(h'+34)  Ioge 
and 

h'+h  =  h'+v(h'+34)  log, 

34 
from  this 


34 
or 

h 

v  = 


34 


520 


PUMPING  MACHINERY 


Given  h  and  //,  v  can  be  found.  This  is  the  amount  of  com- 
pressed air  per  cubic  foot  of  water  pumped.  To  change  this 
to  free  air  per  cubic  foot  of  water  the  following  is  used  : 


_ 

V  a  — 


34 


If  water  is  desired  to  be  delivered  at  a  given  rate  at  the 
bottom  of  the  main  discharge  the  velocity  of  entrance  is  given 
by  the  equation: 


Now 


or 


If  h"  is  found  this  amount  should  be  subtracted  from  h'  to 
get  the  effective  head  on  the  inside.  In  that  case  the  head 
inside  of  the  casing  is  h"  feet  less  than  that  outside  and  this 


FIG.  399. — Compressor  Diagram. 

will  give  w  the  velocity  of  approach.     This  method  gives  the 
rational  manner  of  designing 

To  find  the  power  and  size  of  the  compressor   consider  Fig. 
399,   in   which   the   curve  of  compression    ab   is  of  the  form 


AIR  LIFT  PUMPS  AND  PNEUMATIC  PUMPS 

pvn  =  K.  It  is  desirable  to  have  this  curve  a  rectangular 
hyperbola,  but  that  is  impossible,  as  even  with  water  jackets 
around  the  cylinder  the  exponent  is  rarely  below  1.2  and  1.4 
is  usually  found.  The  work  done  in  compression  is 


Cpdv- 

Jv, 


now 
Hence 


£pdv~K  ('v-n 


_P\V\  -piV'< 

i-n 
The  work  becomes: 

W  =p2V2-{•-— 


n—I 


This  is  the  theoretical  power,  in  which  pi  is  the  suction  pres- 
sure and  FI  the  volume  taken  in  if  there  is  no  clearance  or 
leakage.  There  is  no  effect  of  clearance  on  the  work  of  a  com- 
pressor as  the  work  obtained  on  the  expansion  of  the  clearance 
air  is  equal  to  the  work  required  to  compress  it.  There  is  an 
effect  due  to  leakage  and  this  is  to  change  the  work  by  the 

factor    —  :  —  7^-.     The  volumetric  efficiency  is  the  ratio  of  the 
vol.  eff. 

air  delivered  to  the  amount  which  should  be  delivered.  This  is 
about  95  per  cent.  The  friction  of  the  compressor  increases 
the  amount  of  work  done,  so  that  the  net  H.P.  required  to 
apply  to  the  V\  if  this  is  the  amount  of  free  air  required  per 
minute  is: 


33,000  X  H.P. 


mech.  eff.  vol.  eff.   n  — 


522 


PUMPING   MACHINERY 


The  clearance  effects  the  displacement  by  cutting  out  the 
amount  of  air  cf,  Fig.  400,  and  caring  for  the  amount  fa  only 


FIG.  400. — Compressor  with  Clearance. 

The  factor  which  gives  this  if  cl=  per  cent  clearance  is; 


C/\/>! 


The  displacement  then  is 


D 


from  the  formula 


vol;eff.X{/)' 


FIG.  401. — Multi-stage  Compression. 

for  a  double  acting  compressor  the  size  of  the  cylinder  and 
the  number  of  revolutions  may  be  found  to  give  the  necessary 
free  air  Va.  Allowance  should  be  made  for  piston  rods. 


AIR  LIFT  PUMPS  AND   PNEUMATIC  PUMPS  523 

Since  it  is  not  possible  to  bring  the  curve  of  compression 
close  enough  to  the  isothermal,  a  method  of  multi-staging 
the  compression  has  to  be  devised.  In  this,  Fig.  401,  the  com- 
pression is  carried  to  a  pressure  p2'  when  the  air  is  discharged 
into  a  series  of  tubes  or  around  them,  while  on  the  other  side 
of  the  tube  surface  cold  water  is  circulated.  In  this  manner 
the  air  can  be  left  in  contact  a  sufficient  time  to  be  brought 
to  the  original  temperature,  and  when  taken  into  the  second 
cylinder  at  hf  it  is  on  an  isothermal  from  a.  In  this  way  the 
work  of  the  area  h  W  h'  is  saved.  The  work  in  this  case  is: 


=p2'V2',  since  h'  is  at  the  same  temperature  as  a,  hence 


dW 
This  is  a  minimum  when  -j—f  =  0, 


np  p          n      p2' 

i\~r/:i* 

pj     \pip 


is  the  condition  for  a  minimum  amount  of  work.    Then 


524  PUMPING  MACHINERY 

For  a  three-stage  compressor  the  condition  for  a  minimum  is 


or  in  general  for  m  stages 


p2f"  = 


etc. 


The  Harris  Pneumatic  pump  is  shown  in  Fig.  402.     In 

this  case  to  start  the  pump, 
air  is  sucked  out  of  one  cyl- 
inder A  until  it  is  filled  with 
water,  while  the  pump  line 
on  the  other  side  is  filled 
with  water  and  air  com- 
pressed into  the  cylinder  B 
until  the  water  is  driven 
from  the  chamber.  At  this 
time  the  turn-over  valve  or 
switch  M  is  shifted  so  that 
the  compressed  air  on  one 
side  may  fill  the  pipe  on  the 
other  side.  That  is,  if  the 
volumes  of  the  tanks  are 
each  represented  by  V  and 
the  volume  of  the  pipe  lines 
by  NV  where  N  is  usually  a 
FIG.  402.—  Harris  Pneumatic  Pump.  fraction,  the  air  of  volume 

(i+N)  V  in  the  discharging 

side  at  a  pressure  P0  is  connected  with  a  volume  NV  in  which 
the  pressure  is  called  p0  (when  filled  with  water).    The  temper- 


AIR  LIFT  PUMPS   AND  PNEUMATIC  PUMPS  525 

ature  is  constant  in  these  pipes  since  they  are  exposed  to  the 
atmosphere  and  have  much  surface,  hence  when  the  valve  M 
is  switched 


1+2^ 

After  the  pressure  has  fallen  in  the  system  to  PI  the  com- 
pressor draws  air  from  the  tank  just  discharged  and  builds 
up  the  pressure  in  the  pipe  line  leading  to  the  full  tank.  After 
the  pressure  in  that  tank  becomes  equal  to  the  discharge  head 
P0  the  water  begins  to  discharge.  The  foot  valve  on  the  empty 
tank  does  not  open  until  sufficient  air  has  been  drawn  out  to 
bring  the  pressure  to  pi.  At  that  time  water  just  begins  to 
enter  and  when  the  tank  is  full,  this  pressure  has  become  pQ 
These  two  pressures,  pi  and  p-0,  do  not  differ  by  much  since  the 
head  difference  is  the  height  of  the  tank,  and  this  may  be  very 
small  especially  if  the  tank  is  on  its  side. 

To  find  the  quantity  of  air  to  be  pumped  from  one  tank 
to  another  there  are  three  periods  to  consider. 

ist.  That  in  which  the  pressure  in  the  pipe  line  leading  to  the 
full  tank  is  brought  from  Px  to  P0. 

2d.  That  to  reduce  the  pressure  in  the  suction  tank  to  pi 
while  water  is  discharging  from  the  full  tank. 

3rd.  That  to  fill  the  suction  tank. 

To  find  the  pressure  in  the  suction  tank  at  the  end  of  the 
first  period  the  following  equation  may  be  used: 

P0NV+PnV(i  +N)  =PiV(i  +2N). 

_Pl(i+2N)-P0N 

(i+N) 

Since  N  is  small,  this  quantity  PI  i  will  be  found  to  be  almost 
the  same  as  PI,  but  for  exact  work  the  formula  must  be  used  as 
stated  and  the  displacement  found. 

Suppose  the  displacement  of  the  compressor  is  D  and  this 
volume  is  abstracted  from  the  volume  (i+N)V  of  the  pipe 
and  tank  and  discharged  into  the  volume  NV. 


526  PUMPING  MACHINERY 


The  ratio  of  -      ~      -  is  M,  and  after  the  first  stroke  the 

M 
pressure  PI  is   reduced   to    ^  —  PI,    after     the    second    to 

-  PI\  or  (1r^-\  PI,  and  after  the  t  strokes  the 
Af  +  i  \M  +  i     V         \M  +  i/ 


)  Pj. 


pressure  is 

This,  however,  equals  PH.  hence 


or 

logs  Pi  i- logs  Pi 


tD=Q,  the  quantity  taken  through  the  compressor  during  this 
first  stage.     During  the  second  period  the  compressor  compresses 
against  a  fixed  discharge  pressure  but  the  suction  is  variable. 
In  this  case 


where  dVv  is  the  variation  of  volume  in  the  delivery  tank,  dVc 
that  in  the  compressor,  and  Pj  the  pressure  in  the  tank  being 
emptied  of  air.  Now  at  any  time 

P*V(L  +N)  +P0(VV  +  NV)  =P0V(i  +N)  +p0NV. 
Hence 

PO  __  V(i+N)  _ 

Px     V(i+N)+-NV-Vv-NV 
YQ 

caUing 


Then 

dVc=P~K 


AIR   LIFT   PUMPS  AND    PNEUMATIC   PUMPS  527 

f*vlPt\  CVI     V(i-\-N^ 

Q2  =  Volume  displaced  =    /     — -dVv=    I         .   v      .  .  x — dVv 


r^-dv  =  r  r^+w 
.'-       '- 


Fi  is  the  volume  of  the  discharge  tank  when  the  water  is 
just  entering  the  other  tank,  or  the  pressure  is  pi  in  that  tank, 
hence 


P0Vl  =  V(PG 
=  V[P0 

Now  N  is  a  fraction  and  p0  —  piis  very  small,  hence 

F!=^-(PO-/>I).- 
^o 

Hence 


_ 


Po 

i 


logs 


Neglecting  AT^o  in  comparison  with  P0 


Now,  during  the  third  period,  or  that  of  the  filling  of  the 
suction  chamber,  the  pressure  in  this  chamber  changes  from 


528  PUMPING    MACHINERY 

pi  to  pQ  when  it  is  being  filled  with  water,  and  since  this  change 
is  so  small  the  pressure  may  be  assumed  constant  and  the  quan- 
tity removed  during  this  time  will  be  V.  • 

The  total  displacement  of  the  pump  will  then  be 


In  this,  V  is  the  volume  of  the  water  lifted  by  the  displace- 
ment of  Q  cubic  feet  of  air  by  the  compressor  piston. 

If  the  displacement  of  the  compressor  per  revolution  is 
D  and  it  makes  2V  revolutions  per  minute  the  time  taken  to 
fill  one  tank  is: 

~^  =/'  minutes. 

The  effect  of  friction  of  the  pipes,  according  to  Harris,  is  to 
increase  the  pressure  required  by  the  amount. 

Loss  =  KP  =  —      —  —  -  -  -7  v2P  in  pounds  per  sq.ft. 

/=  length  of  pipe  in  feet; 
d  =  diameter  in  inches; 
v=  velocity  in  ft.  per  sec.; 
P=  pressure  in  pounds  per  sq.ft. 

The  effect  of  this  is  to  multiply  variable  pressures  of  discharge 
by  i+K,  or  what  is  the  same  thing,  to  multiply  the  volume 
which  passes  through  the  compressor  by  i  +K,  and  therefore 
the  time  is  increased  in  the  same  manner. 

The  quantity  K  is  seen  to  vary  as  the  velocity  of  the  air  in 
the  pipe  varies,  but  for  general  use  an  average  value  of  o.i  is 
recommended  by  Harris.  The  quantity  of  air  passing  per 
second  as  the  pressure  changes  is  a  variable,  but  the  velocity 
of  this  air  from  the  suction  tank  remains  about  the  same  because 
its  pressure  changes;  while  that  entering  the  discharge  tank 
varies  with  the  weight,  since  the  pressure  is  constant. 

The  size  of  the  compressor  can  be  determined  when  the 


AIR  LIFT  PUMPS  AND  PNEUMATIC  PUMPS  529 

quantity  of  water  to  be  lifted  per  hour  is  known,  or  the  equation 
for  Q  may  be  used  to  get  the  quantity  of  water  which  may  be 
handled  by  a  given  compressor. 

The  work  done  in  the  compressor  when  Vx  cubic  feet  are 
compressed  has  been  shown  to  be 


n-i 


where  n  is  the  exponent  of  the  compression  curve.  This  quantity 
W  changes  in  value  as  the  air  is  drawn  out.  To  find  when  it  is  a 
maximum  express  it  in  terms  of  the  variable  px,  as  Vx  is  con- 
stant, depending  on  the  displacement  of  the  compressor  while 
P0  is  a  constant, 

nV 
work=—  - 


This  is  a  maximum  when 

d 
dp 


dW  nV  fi      tit*-    «TJE       "1 

==_  _£       n    p       n     _j 

n-ilnr 


For  ^  =  1.4;     px= 
Putting  this  in  for  px  the  work  for  V  cubic  feet  passing 
through  the  compressor  becomes: 


If  the  displacement  of  the  compressor  is  ND  where  D  is 
the  discharge  per  revolution,  the  horse-power  which  must  be 
cared  for  is: 

Max  ,000 

This  is  divided  by  the  efficiency  of  the  compressor  and 
driving  motor  to  get  the  applied  power  at  this  instant. 


UP    AV-'^VV£> 

x.  H.P.  =  1  -I  — 

\nl        33,ooo 


530  PUMPING  MACHINERY 

To  get  the  total  power  employed,  the  following  method 
will  be  used,  for  the  second  stage: 


Now  although  the  compression  within  the  cylinder  is  accord- 
ing to  the  law  pVn=K  the  relation  between  the  pressures  and 
volumes  in  the  tanks  and  pipe  lines  is  isothermal,  so  that  at 
any  time 

px(T.+N)V+Po(Vx+NV)=p9'(L+N)V+PQ(V2+NV), 

or  considering  a   small   amount,   dV  taken  from   the  volume 
V(i  +N)  the  change  becomes 


=  V(i+N)dPx, 


then 
hence 


n^±^P9^  fp^P,-^^n  CdPl 
n—i  J  n-i      J 


The  work  done  during  the  first  period  is 


and  the  work  on  the  last  period  is 

nV  F     £T-i    -        1 
W3=—  -[Pa  "  P1'-p1\. 


AIR  LIFT  PUMPS  AND  PNEUMATIC  PUMPS 


The  useful  work  is 

V(P(}-pi) 

Hence  the  efficiency  of  this  method  of 
pumping  is 


Eff.= 


X 


eff.  compressor  and  motor. 

Harris  has  designed  for  this 
method  of  pumping  a  special  switch 
which  will  reverse  the  turnover  valve 
after  the  compressor  has  made  the 
necessary  number  of  turns  t'N,  or  by 
a  system  of  diaphragms  the  valve  is 
changed  when  the  pressure  pQ  is 
reached.  Leakage  from  the  system 
is  made  up  from  the  atmosphere  and 
allows  a  supply  of  air  when  the  pres- 
sure in  the  suction  is  below  the 
atmosphere  before  the  valve  re- 
verses. If  an  excess  of  air  is  drawn 
in  this  will  do  no  harm,  as  it  will  be 
blown  out  of  the  discharge  before  the 
valve  reverses  the  action. 

It  is  recommended  by  the  build- 
ers of  this  pump  to  make  the  air  pipe 
of  such  a  diameter  that  the  maxi- 
mum velocity  is  not  over  5000  feet 
per  minute  when  it  is  assumed  to 
have  the  volume  of  the  free  air. 
That  is 

Area  air  pipe  = 


/XI4-7XI44X50QO 

The  tanks  are  usually  made  of 
such  a  volume  that  N=±,  or  tanks 
are  four  times  the  vclumes  of  the 


i 


FIG.  403. — Harris  Pneumatic 
Pump. 


532 


PUMPING  MACHINERY 


pipes.     The  discharge  pipes  for  the  water  are  made  2 . 8  diame- 
ter of  air  pipes  and  they  should  discharge  into  a  head  in  the 


FIG.  404. — Separate  Stage  Pump. 

same  manner  as  that  shown  for  the  air  lift  pumps.  Fig.  403 
shows  the  installation  of  a  Harris  pneumatic  pump  for  draining 
the  inverted  syphon  at  High  Bridge.  The  tanks  AB  are  con- 


AIR  LIFT  PUMPS  AND  PNEUMATIC  PUMPS 


533 


nected  to  the  compressor  by  the  pipes  CD  which  extend  to  the 
power-house. 

Fig.  404  illustrates  a  method  of  using  the  system  for  high 
lifts  where  it  is  desired  to  use  an  air  pressure  less  than  that  cor- 
responding to  the  total  lift.  The  air  forced  into  A  lifts  the  water 
against  the  pressure  P0  into  the 
suction  box  B  while  at  this  time 
the  tank  C  is  being  filled.  When 
the  pump  reverses  the  water  in  C 
is  discharged  while  A  is  filled  by 
suction. 

The  method  shown  in  Fig.  405  is 
intended  to  use  compressed  air  under 
a  moderate  pressure  to  raise  water 
a  considerable  height.  Air  is  delivered 
into  B  through  A  at  sufficient  pres- 
sure to  force  water  through  the  foot 
valve  C  into  the  tank  D.  When  the 
air  which  rises  to  the  top  of  B  has 
driven  the  water  from  B  so  that  the 
bottom  of  the  pipe  E  is  exposed,  the 
air  supply  is  cut  off  and  the  air  in 
B  escapes  and  passes  up  to  the  next 
tank  and  lifts  the  water  from  it 
through  the  foot  valve  in  F.  The  air 
which  filled  tank  B  will  fill  tanks  B 
and  D  finally  and  hence  the  pressure 
will  be  half  the  pressure  of  the  supply 
if  the  two  tanks  are  of  equal  size. 
This  means  the  second  lift  can 


FIG.  405.- 


Differential  Air  Lift 
Pump. 


only  be  made  half  of  the  lift  between  the  first  and  second 
tanks.  If  a  third  tank  is  used  the  next  lift  will  be  one  third 
the  original  lift. 

In  this  method  the  air  is  finally  discharged  and  the 
only  use  of  the  internal  energy  of  the  air  is  in  the 
successive  lifts. 

After  the  air  has  been  driven  out  from  the  upper  tank  the 


534  PUMPING   MACHINERY 

lower  tank  is  allowed  to  fill  as  it  is  open  to  the  atmosphere 
through  E  and  D  and  water  will  flow  in  by  gravity. 

There  are  other  compressed  air  pumps  which  act  the  same 
as  these,  but  the  principles  here  brought  out  should  serve  to 
make  their  action  clear. 


CHAPTER  XIV 
CENTRIFUGAL    PUMPS 

THE  last  few  years  have  seen  many  improvements  in  cen- 
trifugal pumps;  much  higher  efficiencies  have  been  obtained 
and  greater  lifts  have  been  overcome.  These  gains  have  resulted 


FIG.  406. — Section  of  Centrifugal  Pump. 

from  a  careful  study  of  the  principles  of  the  pump  and  from 
the  theoretical  design  of  pumps  rather  than  from  the  empirical 
design  which  was  customary  in  earlier  times.  To  understand 
the  theory  of  these  pumps  an  examination  will  be  made  of 
Figs.  406  and  407.  Fig.  406  is  a  longitudinal  section  through 

535 


536  PUMPING  MACHINERY 

a  single-stage  pump,  while  Fig.  407  is  a  cross-section  through 
the  impeller  or  runner  A  and  the  diffuse*  vanes  B.  With  water 
in  the  discharge  tank  D  at  the  level  shown  in  the  figure,  the 
whole  system  will  be  filled  with  water  and  this  water  will  tend 
to  flow  from  the  tank  through  the  pipe  E,  the  volute  casing  C, 


FIG.  407. — Section  through  Impeller,  Diffuser  and  Casing. 

the  diffuser  B,  the  runner  A  and  the  suction  pipe  F.  The 
runner  A  is  mounted  on  the  shaft  H  which  is  supported  by  the 
foot  bearing  I.  If  the  shaft  is  horizontal  two  bearings  take  the 
load  resulting  from  the  weight  of  the  parts. 

If  now  the  shaft  H  be  driven  by  a  belt  on  the  pulley  G  or 
by  a  motor  directly  connected  with  the  shaft,  the  water  is 


CENTRIFUGAL   PUMPS 


537 


driven  by  the  runner  and  forced  to  move  outward  by  the  shape 
of  the  vanes  and  by  centrifugal  force;  moreover  this  motion 
takes  place  in  such  a  manner  that  the  water  leaving  the  outer 
tips  of  the  runner  enters  the  vanes  in  the  diffuser  with  no  shock. 
In  the  diffuser  the  direction  of  motion  is  changed  to  one  more 
nearly  the  direction  to  be  obtained  in  the  volute  chamber  and 
also  the  velocity  is  decreased  so  that  velocity  head  is  changed 
into  pressure  head.  If  the  speed  of  the  runner  is  sufficiently 
great  the  water  will  acquire  enough  velocity  that  when  the 
ends  of  the  vanes  B  are  reached  the  pressure  will  be  greater 
than  that  due  to  the  head  of  the  tank  D. 

As  the  water  is  driven  outward  on  the  wheel  a  partial 


FIG.  408. — Single  and  Double-flow  Impellers. 

vacuum  is  produced  at  the  center  and  water  is  drawn  up  to 
fill  the  space.  At  times  this  water  is  drawn  in  on  two  sides  of 
a  disc  giving  a  double-suction  pump.  The  shaft  is  often  placed 
in  a  horizontal  position,  but  for  simplicity  in  the  derivations 
of  equations  the  shaft  has  been  shown  here  in  a  vertical  position. 
In  some  cases  the  vanes  of  the  diffuser  are  omitted  and  in  this 
event  the  space  around  the  wheel  is  sometimes  called  the 
whirlpool  chamber.  There  may  be  guide  vanes  placed  in  the 
suction  at  the  point  of  entrance  to  the  runner  for  the  purpose 
of  making  the  direction  of  the  water  entering  the  runner  more 
definite.  Fig.  408  illustrates  the  two  forms  of  runner. 

A  study  will  now  be  made  of  the  pressures  in  the  pump 


538  PUMPING  MACHINERY 

and  of  the  action  of  the  pump  in  which  the  following  symbols 
will  be  used: 

r\  =  radius  of  point  of  entrance  = — ; 

r%  =  radius  of  point  of  exit  = — ; 

;= coefficient  of  lost  head  (see  Chapter  V);   l^ 

c=  absolute  velocity  of  water  at  entrance; 

u=  absolute  velocity  of  water  at  exit  from  impeller  or 

entrance  to  diffuser; 

c\  =  velocity  at  entrance  relative  to  impeller; 
c2=  velocity  at  exit  relative  to  impeller; 
Vg=  velocity  in  suction  pipe; 
Vd=  velocity  in  discharge  pipe; 
vv= velocity  in  volute; 
u3  =  velocity  at  exit  from  diffuser; 
v0=  velocity  in  outlet  or  discharge  pipe; 
A  =area  perpendicular  to  c; 
A  with  subscript  =area  perpendicular  to  velocity  with  same 

subscript; 

Wi  =  velocity  of  wheel  or  impeller  at  entrance; 
w2=  velocity  of  impeller  at  exit; 
<*j  =  angular  velocity; 
AT  =  R.P.M.; 

hi  =  suction  head  to  center  of  wheel; 
h2  =  discharge  head  from  center  of  wheel; 
H=hi+h2  =  total  net  head; 

Pa,  pb,  PC,  pd,  etc.,  pressure  at  various  points  in  feet  head; 
w=  weight  of  i  cu.  ft.  of  water; 
a  =  pressure  of  the  atmosphere  in  feet; 
g=  acceleration  of  gravity; 
All  the  above  are  expressed  in  feet  and  pounds. 
At  the  point  of  entrance  into  runner,  the  pressure  in  the 
casing  pa  is  given  by  the  equation, 

Vs2       C2 


CENTRIFUGAL  PUMPS 


539 


When  the  water  enters  the  runner,  the  portion  of  pa  which 
remains  to  overcome  the  friction  of  the  passage  of  the  runner 
and  to  accelerate  the  water  in  so  far  as  the  section  of  this  pas- 
sage may  change  is  called  />&.  This  is  given  by: 


- 


(2) 


—  is  the  pressure  head  due  to  the  action  of  centrifugal 

o 

force.*  This  action,  which  exists  in  all  rotating  bodies  con- 
taining fluids,  is  to  produce  a  hydrostatic  pressure  equal 
to  the  above  expression.  This,  like  any  other  hydrostatic 
pressure,  acts  in  all  directions.  Where  there  is  a  free  surface 
as  in  a  cup,  this  manifests  itself  by  causing  the  water  to  pile 
up  where  the  velocity  is  the  greatest  in  such  a  manner  that  the 
free  surface  is  a  paraboloid  of  revolution.  Where  there  is  no 
free  surface,  there  is  a  force  equal  to  the  height  to  which  the 
water  would  rise  above  the  point  considered  were  a  free  surface 

possible.  This  force  is  equal  to  — — .  The  actual  force  on  a 
rotating  body  which  would  cause  flow  and  overcome  friction 


*  In  Fig.  409,  the  particle  of  water  in 
the  free  surface  of  the  rotating  vessel  at  a 
distance  x  from  the  center  is  moving  with  a 
velocity  xa),  if  o>  is  the  angular  velocity.  It 

w 
is  subject  to  the  centrifugal  force  —  a)2x  ex- 

g 

pressed  in  pounds,  while  the  force  w  acts 
vertically,  due  to  gravity.  The  resultant  of 
these  two  forces  acts  normal  to  the  free  sur- 
face, hence 


dy 
dx 


C.F 

IV 


_    Cx^'X 

Jo    g 


dx= = 


FIG.  409. — Revolving  Vessel. 


540  PUMPINC  MACHINERY 


wx2    w\2 


is  not  the  external  head  acting  at  a  point,  but  that  force  minus 
\2 
g 

,  Ci2       ^vCl2      C22    ,  W22      Wl2  . 

Ate,     *-*+--aF  -2g  V7,'       •     ^) 
Atd,    pa  =  pc~.      .     .     .     .     ./..     .     .     (4) 


(5) 


If  Ws  =Vv  there  is  no  loss  at  discharge  into  the  volute,  but 
if  this  is  not  the  case,  then  there  is  a  sudden  change  in  velocity 
and  there  is  a  loss 


«32 


(6) 

Vv* 

-     (7) 


but'     '      •       p,+fg=h2+a  +      g+-g.    .,i,.i;,;.    (8) 

Substituting  for  pv,  pe,  pd,  pc,  pb,  etc.,  their  values  from  the 
preceding  equations  the  following  results  after  collecting  and 
rearranging: 


V,2      VQ2      U2       W22      C^      C22      Wj        C2 
~r  *•»*          i  ^  --  1  --  T          —  -  —  -  —  -         (  Q  I 

*2g      2g      2g       2g        2g       2g       2g       2g' 

The  passages  of  the  pump  are  completely  filled  with  water, 
hence 

Ac=AiCi=A2c2=AexU=A3u3,  etc.    .     .     .     (10) 

From  these  equations  it  is  seen  that  all  of  the  velocities  may 


CENTRIFUGAL  PUMPS 


541 


be  expressed  in  terms  of  the  areas  and  any  other  velocity.     For 
example: 


2-jy2 


», 


The  Eq.  (9)  now  becomes 


S 


If  there  is  to  be  no  impact  at  entrance  and  exit,  the  velocities 


FIG.  410. — Velocities  at  Entrance — a  positive,  a.^  negative. 


FIG.  411. — Velocities  at  Discharge — «2  negative,  /?  positive. 

at  those  points  must  have  certain  relations  which  are  seen  from 
the  parallelogram  of  velocities.  Calling  the  angle  between 
the  radius  and  c,  a;  between  the  radius  and  c\t  a\\  between 
the  radius  and  c2,  a2]  and  finally  that  between  u  and  the  radius, 
/?;  the  following  relations  may  be  seen  from  Figs.  407,  410  and 
411.  (The  angles  are  taken  as  positive  when  they  are  measured  on 
the  side  of  the  radius  toward  which  the  motion  w^  is  taking 


542  PUMPING  MACHINERY 

place,  provided  the  lines  be  drawn  away  from  the  point  con- 
sidered and  in  direction  of  the  velocity. ) 

Ci  cos  a\  =c  cos  a (14) 

c  sin  a—  Ci  sin  a\  =Wi;  .   ..     .     .     .  (15) 

Ci2=c2+Wi2  —  2cwi  sin  a  or  c2+Wi2—  c12=2cz#1  sin  a;  (16) 

C2  cos  a2  =u  cos  /?;    .' ;   .    '.     '. .    .-    .  (17) 

u  sin  /?—  c2  sin  a2=w2;     »     ...   ,     •  (18) 

c22  =u2  +w22  —  2uw2  sin  /?or  u2  +w22  —c22  =2uw2  sin  /?.  (19) 

Substituting  (16)  and  (18)  in  Eq.  (13)  there  results: 

u2     v02     i  _ 

h— + — =-[uw2sm  3  —  cwism  a].    .     .     (20) 

2g     2g     g 


n 

Calling  St;  —  the  lost  head,  IH,  and  —  the  residual  velocity 

2g  2g 

head,  r#,  the  left-hand  side  of  the  equation  represents  the  total 
head  to  be  accounted  for.  This  when  multiplied  by  one  pound 
is  the  number  of  foot-pounds  of  work  done  on  each  pound  of 
water  passing  from  the  suction  forebay  to  the  discharge  tank. 

#(i+/+r)=work  per  pound  of  water 
=  —  (uw2  sin  p  —  cwi  sin  a). 

o 

To  get  the  total  work  this  expression  must  be  multiplied  by 
the  quantity  of  water  passing  through  the  pump.  All  of  the 
water  taken  into  the  pump  at  a  does  not  leave  the  discharge 
point  of  the  diffuser,  for  a  certain  amount  leaks  past  the  clear- 
ance space  between  the  wheel  and  the  diffuser  or  casing.  The 
quantity  leaking  from  this  point  will  be  computed,  but  for  the 


CENTRIFUGAL  PUMPS  543 

present  it  will  be  called  Qe.     The  amount  desired  is  Q  cubic 
feet  per  second,  and  so  the  amount  Qp,  to  be  pumped  is 

Qp=Q+Qe-,  .......     (21) 

QpwH(i+l+r)=  Work  per  sec.;      ....     (22) 

Qpw-  (uw2  sin  /?  -  cwi  sin  a  )  =Work  per  sec.      (23  ) 

6 

The  expressions  above  are  those  which  give  the  work  required 
to  do  the  hydraulic  work,  but  in  addition  to  this  there  must  be 
energy  supplied  to  overcome  the  resistance  of  the  friction  of  the 
bearings,  stuffing  boxes,  and  the  leakage  water  against  the 
runner.  This  latter  quantity  may  be  quite  large,  while  the 
second  one  depends  on  many  variables.  The  first  may  be 
made  very  small  by  proper  lubrication.  These  will  be  dis- 
cussed later,  but  for  the  present  they  may  be  written  as  fwWu, 
fbWu  and  f8Wu.  The  total  work  then  becomes 


wO 
=——p(uw<>  sin  ff  —  ci&i  sin  a)(i  +f. 

;      (24) 

wQpH(i+l  +  r)(i+fw+fb+f8) 
w(Q+Qe)H(i+l+r)(i+fw+fb+f8) 

The  useful  work  is  wQH  or  in  some  cases,  wQH(i+r),  if 
the  residual  velocity  may  be  utilized  in  any  manner.  The 
efficiency  then  becomes 

wQH  i 

If   ~          TI7  /T^L/7\/T_L.;j^«\/-rl^         1     4         \     -t    \'  '  (25) 


The  efficiency  is  seen  to  vary  with  the  values  of  these  coef- 
ficients; they  should  all  be  as  small  as  possible.  The  deter- 
mination of  these  quantities  will  be  considered  after  a  discus- 
sion of  the  action  of  the  water  in  the  wheel. 

There  are  several  relations  which  may  be  determined  from 


544 


PUMPING  MACHINERY 


Figs.  407,  410  and  411.     Fig.  412  gives  the  two  diagrams  of 
Figs.  409  and  410  in  one. 

a  is  often  made  o°  so  that  the  work  equation  becomes 


now 


(26) 


tl. 

=d  sin 


A,  ' 


c=c 


FIG.  412.  —  Diagrams  at  Entrance  and  Exit. 
Hence  by  combining  these 


eK        . 

kH=—     --  sin/?  sin 
r\Al    g 


or 


u 


2g(kH) 


*     A 

T  2  **•  ex     • 

2 -r-  sin  /?sm 

rlAl 


•     (27) 
(28) 


CENTRIFUGAL  PUMPS  545 

If  it  is  desired  to  express  this  in  terms  of  the  effective  head 
j  Eq.  (20)  may  be  written: 


2' 


•     '     •      (30) 


If  either  of  these  equations  be  used,  the  losses  must  be 
expressed  in  terms  of  H  or  u  before  u  can  be  found.     The  ratios 

^  and  —  may  be  assumed  from  practice  as  well  as  the  ratio 
A  r\ 


For  convenience  it  is  well  to  take  the  components  of  u  and 
c  in  a  radial  direction;  these  will  be  called  ur  and  cr. 

wr  =  wcos/?=c2cos  a:2,  .....     (31) 

cr=c  cos  a=Ci  cos  a\  .....     (32) 

Neumann  uses  the  total  head  kH  in  finding  the  velocities 
w,  ur  and  c2  in  the  following  manner: 

-uw2sm{3=kH,      .     .     .    ^  •  -v    *     (26) 

o 


From  Fig.  412,  Eq.  (18)  and  Eq.  (31), 

ugkH 
w2=u  sin  /?—  c2  sin  a2=  --  wr  tan 


=—-  --  wrtan«2,    .     .     .     .     .....     (33) 


ur  tan 

.     (34) 


546  PUMPING  MACHINERY 

For  a2=o°  this  becomes 


Since         tfr=wcos/?=wsin/?cot/?  =  —      — — ,     .     .     (34^) 

W2 

-  (33)  may  be  written  from  (340): 

2COt/?'  •    •  - (35) 


,    .    .     (36) 


tan 
calling  i.      .    .,,  ,_^-  .  ,,.    (37) 


Eq.  (33)  may  be  used  to  determine  the  angle  «2  in  case  ze>2, 
and  ur  are  known,  and  Eq.  (35)  may  be  used  if  w2  and  /?  are 
known.  Thus 

gkH 


;    (38) 


Ur 

gkH  w£ 

~W2    *-&B 

=~-  -  •  •  •  (39) 


For  the  relative  velocity  c2,  Neumann  uses  the  equations, 

Ur 


C2  = 


COS  a2 


_wcos/?_     u  sin  ft      _         gkH 
2  ~~  cos  «2    .  tan  ft  cos  a2  ~w2  tan  ft  cos  «2*      *     '^° ' 
For  w: 

«--^( 

z£^2  sin  8 

gkH  I     <>kH 

tan 

sin  " 


CENTRIFUGAL  PUMPS  547 

These  equations  are  used  when  certain  quantities  are  known 
or  assumed.  The  assumptions  fix  the  equations  which  would 
be  used.  These  quantities  are  also  obtained  in  another  manner 
by  Neumann.  He  uses  the  expression 


w2=x\//gkH 

and  for  ur  he  proceeds  as  follows: 

cot  8 


(42) 


when  a2=o°,  w2=AvgkH.      .     .     ....    •„-.    .     (43) 


Hence          uro°=\/gkH  cot  pQ°=c2  cos  o°=c2,    .     .     .     (44) 

calling  cot  ft  =  \//2  ^, 



•,    •     ........     .     (45) 


Using  the  value  of  w2  = 

^77  cot  ft    VgkH  cot 


/-r==  ,          ...     (46) 

xVgkH 


and  if  in  general  cot  ft  =  \zX  or  X  =%  c°t2  /?,  the  result  follows: 


VtegkH 
"r  = f— , (47) 

from  Eq.  (41), 

(48) 


The  author  has  re-computed  and  constructs  the  curves 
shown  in  Figs.  413  and  414,  which  are  those  used  by  Neumann. 
By  assuming  various  angles  for  «2  and  ft  the  values  of  u,  w2, 
and  ur  may  be  found  in  terms  of  kH. 

A  series  of  curves  drawn  by  the  use  of  the  equation, 


shows  how  w2  varies  with  different  values  of  ft  and  a2,  but 


548 


PUMPING  MACHINERY 


—  90 


FIG.  413. — Values  of  x.     (After  Neumann.) 


60 


.06  0.10  0.15  0.20  0.25  0.30  0.35  0.40  0.15 

FIG.  414. — Values  of  X.     (After  Neumann.) 


CENTRIFUGAL   PUMPS 


549 


since  w2=x  when  VgkH  =  i,  these  curves  need  not  be  drawn, 
as  the  x  curve  shows  this  variation  to  some  scale. 

It  is  seen  from  an  examination  of  Fig.  413  that  when 
/?=go°,  x  =  i  for  all  values  of  «2  except  90°  and  that  for  0:2=0, 


FIG.  415. — Combined  Entrance  and  Discharge  Diagram. 

all  values  of  x  for  different  values  of  ft  are  unity.  The  varia- 
tion in  x  and  consequently  w2  is  greater  for  changes  in  «2  for 
smaller  values  of  ft  than  for  values  of  ft  near  90°.  When  ft 
has  a  value  of  85°  there  is  little  change,  as  «2  changes  from  50° 
to  -50°.  When  considerable  variation  is  desired  for  different 
values  of  a2>  ft  must  be  taken  smaller,  say  60°. 


550  PUMPING  MACHINERY 

When  a  is  not  o°  the  equations  just  found  will  be  of  differ- 
ent forms.     The  entrance  velocities  will  now  be   considered. 

w\ 
A  new   figure,   using  the  construction  for  w2  =  —r2,    is   given 

in  Fig.   415,   showing  these   other  relations.     In   this   figure, 

A2 

Cl=c2-j-.     Now  cr=c  cos  <x=Ci  cos  «i, 
-*1 

c  sin  a  =cr  tan  a.     .  :  \     .     .  •  '.     .     (49) 
2=w2+ur  tan  a2.       .     .     (50) 


Now  —  (uw2  sin  8—cwi  sin  a)  =kH, 

o 

or  w22+w2ur  tan  a2—  Wicr  tan  a=gkH.   .     .     .     (51) 

Wi=w2 — (52) 

Hence 


r  tan  a     wrtana2l2 


Now  Ur2nr2b-2=cr2xr1bi,  hence 


tan  *  ~~~  tan  "-+g^.    (53) 


In  one  of  these,  cr  must  be  assumed  as  well  as  -     7^,  a2. 

r2    b2 

and  a,  while  in  the  other  cr,  ur,  — ,  a2,  and  a  are  assumed. 

If  the  velocity  u  is  desired  in  terms  of  the  effective  head  H, 
the  following  results; 


CENTRIFUGAL   PUMPS  551 


n/-Cw1  sn  a_-; 

ze?2=wsin/?  —  c2  sin  0:2; 
c2  cos  «2=w  cos  /?; 

w2=u  sin/?—  w  cos  /?  tan  «2=w(sin  /9—  cos  /?  tan  «2);       .     (54) 
ze>i  =c  sin  a  —  c\  sin  «i  =c  (sin  a  —cos  a  tan  a)  ; 


2gH  =  aw2  sin  ^52  —  cos  ft  tan  a2  sin  ft 

[sin  a  -cosa  tan  aj  sin  a  -  S  ;  -       - 


A/  2[sin/?  —  cos/3tana2]  sin/9—  (—  |^)  [sin  a  —  cos  a  tanajsin  a  —  2£—  (—  —  J 

(55) 

and  substituting  the  value  of  w  above,  w2  is  found.     Another 
method  gives  the  following: 


sn     —  WC  sn  a 


u  sin  5  «2  --  +—  cr  tan  a; 
w2      r2 


__> 
~cos/?; 


T 

tan 

Now  z#2  =  u  sin  /?  —  wr  tan  «2, 


gkH    n 
W2  =-  --  \—  cr  tan  a  —  ur  tan 


(56) 


552  PUMPING  MACHINERY 

Substituting  the  value  of  ur  from  Eq.  (56), 

gkH\       tan  0:2!     r\  tana2l 

z02=—  -   i— T   — f-    +-crtana   I— -   -~\.         57 
w2  L        tan  /?  J     r2  L        tan  /?  J 

PI    tan  a  f       tan  a2l 


These  equations  may  be  used  in  determining  the  velocities 
of  various  parts  of  the  wheel  for  a  given  set  of  conditions. 

LOSSES 

The  first  loss  to  be  considered  in  Eq.  (9)  is  the  loss  in  the 
suction  pipe  ^,.  This  coefficient  includes  the  losses  due  to 
entrance,  friction  in  pipe  and  bends.  These  losses  have  been 
discussed  in  Chapter  V,  and  from  what  has  been  given  there 
the  following  may  be  written: 


The  equation  shows  that    ^s  varies   directly   with   /   and 
inversely  with  d,  and  moreover  the  total  loss  varies  with  v82. 

It  is  an  important  matter  to  determine  just  how  much  ^  — 
may  be.     In  any  case 

p^a-h.-^-^pt,      .:.    .    .     (60) 

where  pt  in  feet  is  the  steam  pressure  corresponding  to  the  tem- 
perature of  the  water  in  the  suction  pipe.  If  this  inequality  is  not 
true,  the  suction  column  will  part.  Of  course,  the  term  hi  is 
the  one  which  is  usually  responsible  for  the  inequality  not 
holding,  and  a  change  in  it  may  make  this  true.  It  may  be, 


CENTRIFUGAL   PUMPS  553 

though,  that  vs2  is  so  great  and  that  -7  is  so  great  that   the 

v  2 
term  £,—  is  responsible  for  this  inequality   not  holding.     In 

such  a  case  the  diameter  of  the  suction  must  be  increased, 

I 

giving  a  smaller  value  of  -7  and  vs2.     The  bends  in  the  pipe 

must  be  gradual  to  keep  the  value  of  m  as  small  as  possible. 
Although  the  velocity  in  the  suction  pipe  should  be  such 
that  there  is  no  danger  of  the  column  of  water  parting,  the  loss 
should  be  reduced  to  a  low  value.  It  is  difficult  to  tell  just 
what  velocity  to  allow  to  give  the  loss  in  the  suction  pipe  a 
value  consistent  with  the  other  losses.  A  method  used  by  some 
is  to  make  the  velocity  in  the  suction  pipe  equal  to  the  velocity 
caused  by  a  certain  percentage  of  the  total  head.  Neumann 
uses  2  per  cent  of  the  head  kH.  If  the  suction  pipe  is  short 
and  direct,  this  value  will  be  used  while  with  a  long  pipe  (say 
50  diameters)  a  i  per  cent  value  will  be  employed.  This  gives 

v8  =  8. 02*/-—  (short  pipe),     .     .     .     .     .     (61) 

'    +J 

8  02      

=-^-VkH  (long  pipe) (62) 

In  any  case  a  discussion  similar  to  that  given  under  the  discharge 
pipe  could  be  used  at  this  point. 

c2 
The  loss    ^ent—  is   due  to  the  incorrect  angular  relation 

between  the  fixed  vanes  (if  used)  and  the  movable  ones  at 
entrance  and  also  to  the  interference  between  the  moving  or 
fixed  vanes  with  the  flow  of  water,  causing  impact.  The  same 

u2 
kind  of  losses  occur  at  exit  in  the  term  i;ex — .     For  this  reason 

these  two  will  be  considered  at  the  same  time. 

In  Fig.  416  a  series  of  movable  vanes  is  placed  opposite 
a  set  of  fixed  vanes.  If  the  velocities  are  as  shown  in  the 
diagram,  the  water  as  it  leaves  the  moving  vanes  is  traveling 


554 


PUMPING  MACHINERY 


through  space  as  shown  in  the  figure.  When  this  water  strikes 
the  blunt  vane,  there  is  a  certain  amount  of  loss  due  to  sudden 
contraction,  and  moreover  when  the  water  leaves  the  blunt  end 
of  the  moving  vane  there  is  a  loss  due  to  sudden  enlargement. 


Diffuser 


FIG.  416. — Interference  of  Blunt  Vanes. 

To  cut  down  these  losses  the  edges  of  the  vanes  are  sharpened 
as  shown  in  Fig.  417,  and  there  should  be  a  certain  amount  of 
clearance  between  the  two  sets  of  vanes.  This  clearance  permits 
the  'water  leaving  one  set  of  vanes  to  come  together  as  one 


FIG.  417. — Pointed  Vanes. 

mass  of  water  before  entering  the  other  set  of  vanes.  The 
action  of  one  set  of  vanes  on  the  other,  if  the  vanes  are  so  close 
that  there  is  a  space  between  the  masses  of  water  discharging 
from  each  channel,  should  be  such  that  much  noise  and  shock 
would  result. 


CENTRIFUGAL  PUMPS 


555 


By  cutting  the  edges  of  the  vanes  and  by  separating  the 
various  sets,  these  coefficients  ^ent  and  ^ex  are  made  small. 

Now  these  losses  are  those  which  occur  when  the  relations 
between  the  various  velocities  are  such  that  there  is  no  impact. 
In  many  pumps,  however,  there  is  a  change  in  the  pressure  or 
quantity  discharged  while  the  pump  is  driven  at  a  constant 
speed.  This  means  a  change  in  the  velocities  c2  and  u  on  the 
outlet  side,  since  these  are  given  by  the  equations: 


If  Q  is  increased  c2  and  u  will  be  increased,  while  a  decrease 


C2  (Decreased  quantity) 


FIG.  418. — Effect  of  Changing  Quantit}*-  with  Fixed  Speed. 

in  Q  will  change  c2  and  u  in  the  opposite  manner.  These 
changes  are  shown  in  Fig.  418.  It  is  seen  that  the  resultant 
of  c2  and  w2  is  not  in  the  direction  of  u  nor  of  the  value  of  u. 
For  this  reason  there  is  a  loss  due  to  a  sudden  change  in  velocity 
of 


—  1/^2 


=loss, 


and  a  loss  due  to  the  impact  against  the  side  of  the  vane  which 
may  not  be  very  large,  as  has  been  proven  by  experiments  on 
the  loss  due  to  the  impact  of  a  jet  against  a  flat  plate  at  right 
angles  to  it.  A  similar  loss  to  this  occurs  at  entrance  when 
the  quantity  changes  from  the  normal  amount. 


556  PUMPING    MACHINERY 

The  losses  in  the  vanes,  C»~~i  and  that  in  the  diffuser,^  —  , 

are  similar  to  loss  in  pipes  and  bends.     They  depend  on  the 

£ 
length  of   the  passage  /,  the  hydraulic  radius  *—  (A  =area   of 

passage,  />=the  perimeter  of  passage),  the  smoothness  of  the 
passage,  and  its  radius  of  curvature.  These  losses  are  of  the 
form 

S.-fcfc'j. 
P 

k,  being  the  coefficient  for  straight  pipes,  =:  —  =0.005,  and  k 

4 
being  the  coefficient  for  curvature. 

The  loss  —  -  --  —  on  entrance  to  the  volute  casing  around 

2# 
the  diffuser  may  be  eliminated  by  making  v0=u3.     In  any 

case  it  is  a  small  quantity. 

o 

The  loss  ^  —  is  the  loss  which  occurs  in  the  discharge  pipe 

and  is,  therefore,  similar  to  the  loss  in  the  discharge  pipe  of 
any  pump, 


In  the  design  of  an  installation  a  number  of  pipes  should 
be  investigated  for  loss  and  for  cost.  That  one  should  be  taken 
which  gives  a  result  such  that  the  use  of  a  larger  pipe  would 
increase  the  yearly  cost  of  interest,  depreciation,  insurance 
and  taxes  on  the  investment,  more  than  the  saving  in  the 
decreased  cost  of  power  by  the  reduction  in  the  lost  head, 
while  the  use  of  a  smaller  pipe  would,  on  account  of  the  greater 
loss  in  head,  increase  the  cost  of  power  more  than  the  saving 
in  the  yearly  cost  for  interest,  depreciation,  insurance  and  taxes. 
A  velocity  may  be  assumed  in  terms  of  the  total  head  kH  as 
in  the  case  of  the  suction  pipe,  using  the  same  values;  however, 
the  method  given  above  is  the  better  one. 


CENTRIFUGAL   PUMPS  557 

When  time  will  not  permit  of  this  investigation,  the  same 
values  may  be  used  as  for  the  suction,  giving 


Q 

^0=8.02--, 


or, 


In  this  work  for  design  £/  and  V  may  each  be  taken  as 
:—  —  -,  and  these  terms  will  include  the  losses  at  entrance  to 


nmner  and  to  diffuser. 


LEAKAGE 


The  leakage  of  water  through  the  clearance  space  is  due  to 
the  difference  in  pressure  within  the  runner  and  on  the  outside. 
Fig.  419  shows  several  forms  of  impellers.  The  pressure  on  the 
inside  of  the  runner  is  pc  (Fig.  406),  while  on  the  outside,  the 
pressure  is  pa.  If  the  clearance  is  t  feet  and  the  circumference 
is  27iT2,  the  area  through  which  leakage  may  occur  is 

2  X  2xr2t. 
The  velocity  is 


Vi=cv8.o2\/pc—pa,      ......     (63) 


cv=  coefficient  of  velocity. 
FromEqs.  (i)to(8)," 


u2        u2       vi2  v02    v02 


W     c2 

----;       pe=pt, 

U2       Vn2       C2         U2 

+  ^  +  ^  +  ^+^}~+---.    .     (64) 


c2  u2 

Neumann  proposes  to  assume  —  =  (  Cent  +  V)  —  . 

o  o 


558 


PUMPING   MACHINERY 


CENTRIFUGAL  PUMPS 


559 


Then 
Hence 


Qe=cv47tr2t  8.02*  kH 


(66) 


From  this  it  is  seen  that  the  leakage  is  a  function  of  the  total 
head,  but  it  decreases  with  an  increase  in  the  absolute  velocity 


FIG.  420. — Non-aligning  Ring  Oiling  Bearing. 

of  discharge  from  the  pump.  If  u  is  small,  the  velocity  of  dis- 
charge through  the  clearance  space  is  practically  that  due  to  the 
total  head  acting  on  the  pump.  For  this  reason  it  is  necessary  to 
make  the  clearance  between  the  edge  of  the  runner  and  the  casing 
as  small  as  possible.  At  times  the  leakage  is  diminished  by 
placing  rings  on  the  casing  and  runner  as  at  6,  Fig.  419.  This 
increases  the  pressure  in  the  space  a  to  pa'  and  so  decreases 
the  flow.  Such  an  arrangement  will  increase  the  end  thrust  if 
the  runner  is  not  a  double  one. 

Friction  at  Bearings  and  Stuffing  Boxes.  The  friction  at 
a  bearing  may  be  made  very  small  by  the  use  of  ring  oiling 
bearings  and  by  the  use  of  spherical  pivoted  bearing  boxes. 
Such  boxes  are  shown  in  Figs.  420  and  421.  With  the  proper 


560 


PUMPING   MACHINERY 


lubrication  the  coefficient  of  frictioji  will  be  between  o.oi  and 
0.004,  so  tnat  tne  work  of  friction  will  be 


.....     (67) 

W=  weight  of  impeller  or  shaft  for  a  motor-driven  pump; 
=  resultant  of  weight  and  belt  pull  for  a  belt-driven  pump; 
r=  radius  of  shaft  in  feet; 
F  =  foot-pounds  of  work  per  minute  due  to  bearing  friction. 

The  coefficient  /JL  for  lubricated  journals    varies   with   the 
pressure  and  with  the  velocity.     This  must  be  taken  into  account 


FIG.  421. — Self-aligning  Ring  Oiling  Bearings. 

in  the  determination  of  the  friction.  Fig.  422  shows  the  vari- 
ation of  JJL  with  pa,  the  pressure  per  square  inch  of  projected 
area  and  with  velocity.  These  curves  are  for  a  good  grade  of 
spindle  oil. 

In  fixing  the  length  of  the  bearing,  the  designer  would 
assume  the  length  from  practice,  as  the  allowance  of  50  to  200 
Ibs.  per  square  inch  of  projected  area  would  not  give  a  bearing 
long  enough  for  a  motor-driven  pump.  In  the  belt-driven 
pump  the  length  is  determined  by  the  formula  below,  using  50 
Ibs.  as  the  allowable  bearing  pressure, 


(68) 


CENTRIFUGAL  PUMPS 


561 


This  friction  may  be  increased  if  the  bearings  bind  in  any 
way,  and  for  that  reason  self-aligning  bearings  should  be  used. 
The  plain  ring  oiling  bearing  shown  in  Fig.  420  is  often  used. 
With  the  bearings  in  good  condition  the  friction  at  these  points 
will  be  a  very  small  part  of  the  total  power,  especially  with 
motor-driven  pumps.  The  end  thrust  due  to  the  pressure  action 
of  the  entering  water  in  a  single-flow  pump  is  carried  by  a 
thrust  collar  bearing  as, shown  in  Fig.  423.  The  pressure  is 


0.04 


0.03 


0.02 


0.01 


Values  of  Ibs.  per  Sq.  in.  or  Ft.  per  Sec. 


50          100  Ibs.       150 
10'ft. 


200 

£)' 


250 


300     -       '6M 
30  ' 


400 
40' 


450 


FIG.  422. — Curves  of  Variation  of  Coefficient  of  Friction. 

represented  by  P0  and  the  necessary  area  is  found  by  allowing 
a  bearing  pressure  of  60  Ibs.  per  square  inch, 


(69) 


The  friction  from  this  source  is  found  by  using  a  coefficient 
of  friction  /*  which  has  the  value  0.035  for  properly  lubricated 
thrust  bearings: 


F.  = 


N 


(70) 


F7/=work  of  friction  per  minute  in  foot-pounds; 
di  =  outer  diameter  of  collars  in  inches; 
d= diameter  of  shaft  in  inches. 


562  PUMPING   MACHINERY 

The  value  of  /*  for  thrust  bearings  does  not  vary  much  with 
the  speed  or  pressure.  The  value  of  p  given  above  is  a  mean 
value  determined  by  Tower  in  his  noted  experiments. 

The  friction  from  stuffing  boxes  depends  on  the  amount, 
kind  and  arrangement  of  the  packing.  From  experiments 
performed  by  Professor  Benjamin,  the  formula  below  may  be  used 
as  a  guide  for  the  friction  to  be  expected  from  each  stuffing 
box  when  the  nuts  are  tightened  by  the  application  of  16  pounds 
at  the  end  of  a  7-inch  wrench.  (The  bolts  are  spaced  6  inches 
apart  on  the  pitch  circle.) 


1=  length  of  packing  in  feet; 

v=  velocity  of  rubbing  in  feet  per  minute. 

To  cut  down  the  amount  of  friction  from  the  stuffing  boxes, 
wrater  is  sometimes  allowed  to  enter  the  space  around  the  shaft 
on  the  suction  side  as  shown  in  Fig.  430,  so  that  air  is  not 
allowed  to  enter  the  suction  space.  This  water  may  be  drawn 
in  without  interfering  with  the  action  of  the  pump  and  at  the 
same  time  the  stuffing  box  at  the  end  of  the  shaft  does  not  have 
to  be  tight  enough  to  be  air  tight.  The  stuffing  box  on  the  other 
side  of  the  pump  is  under  high  pressure  and  water  tends  to 
flow  outward.  This  should  be  allowed  to  happen,  as  the  small 
amount  of  leakage  tends  to  cool  the  stuffing  box.  In  operating 
pumps  the  stuffing  boxes  on  each  side  should  be  allowed  to 
drip,  as  this  not  only  cools  the  shaft  but  indicates  that  the  pack- 
ing is  not  too  tight. 

Friction  of  Water  on  Back  of  Impeller.  The  friction  of 
water  on  the  back  of  the  impeller  is  an  important  item  and 
may  amount  to  a  large  part  of  the  loss  in  the  pump. 

Let  r  be  the  radius  to  a  particular  part  of  the  impeller  and 
2V  the  revolutions  per  minute.  If  ^>=the  friction  on  unit  area 
at  this  point,  the  resistance  of  an  elementary  ring  is 


Now,  p  is  proportional  to  the  square  of  the  relative  velocity 


CENTRIFUGAL   PUMPS  563 

between  the  water  and  the  surface  of  the  impeller,  and  since 
this  would  vary  from  the  center  to  the  outside, 

p=k(2nNr)2, 
hence, 

/Vo 

Total     W/t/  =  (27r)4&V3  I     r*dr 

J'c  .      ...    (71) 


Professor  F.  G.  Hesse  performed  a  series  of  experiments  at 
the  University  of  California  on  a  step  bearing  in  which  the 
friction  is  about  the  same  as  that  occurring  on  the  two  sides 
of  an  impeller,  and  derives  the  following  formula: 

T^  =  i76Xio-W3Z)5ft.lb.  per  minute,    .    .     .     (72) 
from  this  &'=58Xio-5. 

D=  diameter  in  feet; 

Stodola  gives  the  loss  from  the  rotation  of  steam  turbine  discs 
in  steam  as 

/  \  ^ 

N'  =0.0721  Di2' 

where  N'=HP  due  to  friction; 
DI  =  diameter  in  feet; 

HI  =  peripheral  velocity  of  disc  in  feet  per  second; 
f=  weight  of  i  cu.ft.  of  medium. 

He  gives  this  as  an  empirical  formula  from  the  results  of  Odell 
and  from  his  own  experiments  as  well  as  the  experimental  work 
of  Lewecki. 

If  this  is  changed  to  a  form  similar  to  that  given  above  with 
7—62.5  it  reduces  to  work  per  minute, 


....     (73) 

Of  course  this  is  not  intended  to  be  used  with  water  as  the 
medium  in  which  the  disc  is  revolving,  but  the  form  is  quite 
similar  and  the  formula  is  given  here  for  reference. 


564 


PUMPING  MACHINERY 


Unwin  in  his  hydraulics  describes  experiments  for  the  deter- 
mination of  this  friction  and  reduces  the  following  equation: 


,   ;.     .      (74) 


or 


FIG.  423. — Sectional  View  of  Worthington  Volute  Pump. 

In  order  to  cut  down  this  friction  loss  the  leakage  water 
may  be  drawn  off  so  that  it  will  not  completely  fill  the  space 
between  the  impeller  and  the  casing.  This,  however,  means 
increased  leakage,  as  the  pressure  between  the  wheel  and  the 
space  around  it  at  the  discharge  edge  will  be  greater  in  this 
case.  At  times  air  under  pressure  is  put  in  this-  space  to  keep 
the  water  back  and  thus  reduce  friction.  It  is  well  to  remem- 


CENTRIFUGAL   PUMPS  565 

ber  that  since  the  friction  varies  as  N3D5  it  is  more  important 
to  use  a  small  diameter  and  large  N  than  a  large  D  and  small  N 
when  the  value  xD2N  =w%  is  fixed.  This  fact  appears  in  practice 
where  small  impellers  are  found. 

FORMS  OF  CENTRIFUGAL  PUMPS 

There  are  many  forms  of  centrifugal  pumps.  When  the 
guide  vanes  are  omitted  the  water  passes  directly  into  the 
volute  casing  through  a  filling  -ring.  This  was  the  original  form 
of  centrifugal  pump.  By  some,  manufacturers  it  is  called  a 
volute  centrifugal  pump.  Fig.  423  shows  the  section  through 
a  single -suction  Worthington  Standard  volute  pump.  Fig.  424 
shows  a  double-flow  volute  pump  of  one  of  the  earlier  forms.  In 
Fig.  423,  water  enters  the  suction  head  A  at  the  point  B  where  the 
suction)  pipe  is  attached.  From  the  head  it  enters  the  impeller 
C,  meeting  the  vanes  D,  which  force  the  water  outward  into  the 
filling  ring  E  and  finally  into  the  volute  casing  F.  The  impeller 
is  of  the  form  shown :  in  Fig.  408.  The  vanes  are  enclosed  on 
each  side  by  a  disc  so  that  the  water  does  not  come  in  contact 
with  the  bearing  head  G  on  one  side  or  the  suction  head  H  on 
the  other.  The  form  of  impeller  in  Fig.  424  is  not  of  the  enclosed 
type  and  consequently  there  is  considerable  water  friction. 
Moreover,  the  mechanical  friction  is  greater  in  the  unenclosed 
type,  and  as  the  vanes  fit  so  closely  to  the  heads  a  slight  amount 
of  end  play  will  cause  rubbing. 

In  the  impeller  there  is  an  increase  of  energy  due  to  the 
increase  in  the  peripheral  speed.  This  increase  of  energy  may 
take  the  form  of  increase  in  pressure  if  u  is  the  same  as  c,  or 
it  may  be  in  the  form  of  increase  in  kinetic  energy  if  pa  =pd- 
In  general  there  is  an  increase  in  both  the  kinetic  energy  arid 
the  potential  energy  or  pressure.  In  general  pa<pd,  so  that 
there  is  leakage  from  the  outlet  of  the  impeller  into  the  space 
around  this  part  of  the  pump. 

The  bushing  rings  /,  /,  prevent  excessive  leakage  from  the 
space  around  the  impeller  into  the  suction  space.  The  space 
within  the  right-hand  bushing  ring  is  connected  to  the  suction 


566 


PUMPING  MACHINERY 


space  by  the  openings  in  the  central  part  of  the  impeller  disc, 
so  that  any  leakage  at  the  ring  will  be  taken  into  the  suction 
space.  These  openings  also  equalize  the  pressure  at  the  center 
of  the  wheel  and  reduce  the  end  thrust.  The  stuffing  boxes 
at  the  two  sides  of  the  pump  are  equipped  with  a  cage  J  between 
the  two  sets  of  packing  rings  for  the  introduction  of 'oil  or  grease 
from  cups  attached  to  the  stuffing  boxes  K  and  to  connect  water 
pipes  from  discharge  casing  to  farm  a  water  seal.  The  outboard 
bearing  L  and  the  suction  bearing  M  are  provided  with  oil  rings 


FIG.  424. — Double  Flow  Unincased  Impeller. 

N  and  the  oil  cellar  0.  The  oil  is  introduced  through  the  opening 
capped  by  the  cover  P.  The  thrust  bearing  Q,  supporting  the 
small  end  thrust,  is  principally  used  to  keep  the  impeller  in 
a  central  position.  The  cap  R  on  the  end  of  the  shaft  is  an  oil 
thrower.  The  shaft  S  is  supported  by  two  bearings  and  turned 
with  a  small  shoulder  at  the  center  so  that  the  impeller  may 
be  secured  in  a  given  longitudinal  position  by  the  nut.  The 
impeller  is  keyed  to  the  shaft  and  the  coupling  used  to  attach 
this  to  the  motor  is  also  keyed.  The  ribs  around  the  volute 
casing  are  used  to  strengthen  that  casting,  which  is  not  a  com- 


CENTRIFUGAL   PUMPS 


567 


plete  pipe  on  account  of  the  opening  from  the  filling  ring,  and 
for  that  reason  the  bending  action  on  the  casting  requires  addi- 
tional ribs. 

It  is  the  object  of  the  whirlpool  chamber  to  reduce  the 
absolute  velocity  of  discharge,  changing  it  into  pressure.  It 
does  not  accomplish  this  efficiently,  as  there  is  shock  here  and 
a  sudden  change  in  velocity  which  means  a  loss.  By  having 
this  chamber  gradually  increase  in  area  the  velocity  of  the 
water  decreases  as  it  passes  through,  and  with  this  action  the 


FIG.  425. — Worthington  Belted  Volute  Pump. 

pressure  increases.  To  make  this  change  in  an  effective  manner, 
however,  the  diffuser  is  used,  in  which  there  is  no  shock,  and  the 
velocity  head  is  changed  into  pressure. 

Fig.  425  illustrates  one  of  the  Worthington  single-flow 
volute  pumps  for  a  belt  drive,  while  Fig.  426  illustrates  a  double- 
flow  volute  pump,  motor  driven.  These  figures  show  the 
external  appearance  of  the  pumps;  the  pipes  for  water  cir- 
culation in  the  stuffing  boxes  and  thrust  bearing;  the  stiffen- 
ing webs;  the  method  of  supporting  the  casing  and  the  method 
of  attaching  the  heads  to  the  casing.  The  double-flow  pump 


568  PUMPING   MACHINERY 

of  Fig.  426  is  not  equipped  with  a  thrust  bearing,  as  in  this  type 
of  pump  there  should  be  no  end  thrust. 

These  pumps  are  intended  primarily  for  heads  up  to  65 
feet,  and  in  most  cases  when  this  head  is  exceeded  the  turbine 
type  of  pump  with  a  set  of  diffuser  vanes  would  be  used.  The 
impellers  built  by  Worthington  are  specially  designed  to  suit 
the  conditions  of  the  particular  service  required  by  the  pur- 
chaser, and  this  should  always  be  done,  as  each  speed  and  head 


FIG.  426. — Motor  Driven  Worthington  Double  Flow  Volute  Pump. 

requires  a  definite  form  of  impeller.  The  suction  head  shown 
in  Fig.  423  is  so  arranged  that  water  may  enter  from  all  parts, 
giving  an  adequate  supply  to  the  impeller. 

These  pumps  are  built  in  different  sizes,  as  shown  by  the 
tables  below;  the  first  table  giving  the  capacities,  speeds,  heads 
and  power;  the  second,  the  dimensions  of  the  casing. 

The  pumps  shown  so  far  have  been  arranged  with  the  suc- 
tion below,  while  the  discharge  has  been  taken  off  vertically. 
This  arrangement  may  be  changed  to  suit  as  shown  in  Fig. 
427,  which  illustrates  a  few  of  the  different  arrangements  pos- 
sible with  their  designating  numbers. 


CENTRIFUGAL   PUMPS 


569 


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PUMPING    MACHINERY 


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572  PUMPING  MACHINERY 

A  recent  design  of  volute  pump  for  a  low  head  but  large 
capacity  is  shown  in  Fig.  428.  This  pump  is  known  as  a  tri- 
rotor  volute  pump.  It  was  installed  in  the  Interborough  Rapid 
Transit  Station  for  circulating  the  water  through  the  surface 
condensers.  It  is  in  reality  three  double-flow  volute  pumps 
with  four  vertical  suction  pipes  rising  to  the  center  of  the  casing 
and  threading  among  the  three  horizontal  discharge  pipes 
which  unite  and  discharge  from  one  large  opening  at  the  front 
of  the  bed  plate.  The  head  being  small,  there  are  no  stiffen- 
ing ribs  found  on  the  exterior  of  the  casing.  The  ring  oil 
bearings,  thrust  bearing,  water-supply  pipes  for  stuffing  boxes 


FIG.  428. — Worthington  Turbine  Driven  Tri-rotor  Volute  Pump. 

and  thrust  bearing  as  well  as  the  general  arrangements  for  casting 
the  casing  and  heads  are  clearly  seen.  The  figure  also  illustrates 
the  method  of  driving  the  pump  by  a  steam  turbine. 

The  section,  Fig.  429,  is  given  not  only  to  show  the  method 
of  bringing  the  water  to  the  various  suction  chambers  and  taking 
it  from  the  separate  volute  casings  into  the  one  discharge 
opening,  but  the  arrangement  of  the  impeller  is  to  be  noted. 
The  impellers  here  are  the  typical  double-flow  volute  runners 
and  the  bushing  rings,  filling  sleeves  and  bearings  are  all  in 
evidence.  These  two  figures  illustrate  the  type  of  split-casing 
pumps,  which  type  is  usually  employed  in  sugar  factories,  or 
gas  plants  where  the  pump  may  require  cleaning  frequently. 


HI 


FIG.  429.— Worthington  Tri-i 


c  r  Volute  Pump. 


(To  face  page  572} 


CENTRIFUGAL  PUMPS 


573 


When  higher  heads  than  60  or  70  feet  are  to  be  overcome 
the  volute  pump  is  replaced  by  the  form  of  pump  used  in  the 
development  of  the  theory  of  centrifugal  pumps,  one  with 
diffuser  vanes.  This  type  of  pump  is  known  as  a  turbine  pump. 
The  rotative  speed  of  these  pumps  is  often  quite  high,  as  the 
diameter  of  the  impeller  may  then  be  made  small.  The  dis- 
charge velocity  in  these  pumps  is  quite  high  and  hence  to  change 
this  over  to  pressure  head  efficiently  the  diffuser  is  used.  At 
times  the  head  is  so  great  that  the  losses  would  be  large,  and 

DISCHARGE 


SUCTION 
FIG.  430. — Section  of  Worthington  Standard  Turbine  Pump. 

in  such  a  case  two  pumps  would  be  placed  in  a  series,  the  first 
one  caring  for  part  of  the  pressure  and  the  second  for  the  remain- 
ing part.  In  some  cases  as  many  as  ten  steps  are  used.  These 
separate  pumps  may  be  combined  in  one  casing,  giving  a  multi- 
stage turbine  pump,  each  stage  caring  for  a  head  of  from  75 
to  200  feet.  Fig.  430  is  a  section  of  a  Wortliington  two-stage 
pump.  The  action  of  this  pump  is  similar  to  the  one  described 
earlier,  except  that  as  the  water  leaves  the  impeller  A  it  enters 
the  diffuser  B,  which  has  channels  formed  as  shown  in  Fig. 
431.  The  diffuser  is  made  up  of  two  parts,  the  diffusion  ring 
C  with  vanes  and  the  diffusion  ring  without  vanes  D,  In  this 


574 


PUMPING   MACHINERY 


the  velocity  is  reduced  so  that  water  passes  through  the  channel 
ring  E  with  a  low  velocity  but  with  considerable  pressure, 
entering  the  second  impeller  F,  which  discharges  into  a  second 


FIG.  431. — Diffuser. 

diffusion  ring  from  which  it  discharges  into  the  final  discharge 
channel  G  in  the  casing  H  of  the  pump,  leaving  the  dis- 
charge /.  The  casing  is  cast  with  the  outboard  head  solid  and 
it  is  machined  on  the  inside  so  that  the  diffusion  rings  and 


FlG.  432. — Four  Impellers  and  Shaft. 

channel  rings  can  be  introduced  from  one  side  with  the  suction 
head  bolting  into  position  on  account  of  the  turned  projection 
on  its  inner  face.  All  parts  are  thus  held  in  their  proper  posi- 
tions. Such  a  construction  makes  it  possible  to  dismantle 
the  pump  readily  for  examination  and  repair  and  yet  reassemble 


CENTRIFUGAL  PUMPS  575 

with  all  parts  coming  into  their  proper  positions.  From  Fig.  431 
the  manner  of  forming  the  diffusion  ring  with  vanes  is  seen  and 
it  is  evident  that  it  is  possible  to  polish  these  smooth  to  reduce 
the  friction  losses.  Fig.  432  shows  the  four  impellers  of  a  four- 
stage  pump.  The  entrance  channels  are  seen  on  the  right-hand 
sides  of  the  discs.  At  the  end  of  the  shaft  are  seen  the  collars 
of  the  thrust  bearing.  In  Fig.  430  the  stuffing  boxes  are  shown 
attached  to  the  heads  by  flanges  and  the  bearings  are  seen  on 
heavy  brackets.  The  thrust  bearing  is  made 
in  halves  and  bolted  to  the  end  of  the  sue- 
tion  bearing.  This  construction,  shown  in 
Fig.  433,  is  necessary  for  the  introduction  of 
the  shaft.  The  cooling  of  this  bearing  is 

'          FIG.  433.— Worthmgton 

accomplished  by  a  current  of  water  through        Thrust  Bearing, 
the  shell,  while   continuous   lubrication    is 
effected  by  the  oil   thrower  /.      The   manner   of  introducing 
water  into  the  suction  stuffing  box  is  evident  from  the  figure. 
The  bushing  rings  at  K  prevent  the  excessive  leakage  of  water 
into  the  suction. 

The  method  of  filling  the  space  between  two  impellers  by 
means  of  the  distance  bushing  M  and  the  collar  N  of  the  chan- 
nel ring  may  be  seen  in  Fig.  430.  The  construction  of  a  stage 
pump  of  greater  number  of  stages  would  be  similar  to  this. 

The  materials  used  for  the  various  parts  of  the  pump  depend 
on  the  kind  of  liquid  handled.  The  impellers  are  of  cast  iron 
or  bronze  and  are  finished  smooth  on  the  outside  and  inside. 
They  are  balanced.  The  diffusion  rings  are  made  of  the  same 
materials  as  the  impeller.  The  shaft  is  made  of  steel  or  Tobin 
bronze. 

Fig.  434  illustrates  the  method  of  constructing  a  two-stage 
pump  as  made  by  the  Alberger  Pump  Company.  The  arrange- 
ment of  the  diffusing  vanes,  the  channel  ring,  the  bearings, 
the  stuffing  boxes  and  the  method  of  uniting  the  various  parts 
together  are  all  evident  from  the  figure.  The  pumps  built  by 
this  company  are  also  designed  for  the  actual  conditions  of 
operation  and  the  surfaces  of  the  impeller  and  diffuser  vanes  are 
highly  finished  all  over  to  reduce  the  losses. 


576 


PUMPING   MACHINERY 


CENTRIFUGAL   PUMPS 


577 


Fig.  435  illustrates  a  ten-stage  motor-driven  pump  while 
Fig.  436  is  an  eight-stage  pump.  These  pumps  are  for  high 
heads. 

When  high  heads  are  to  be  overcome  large  diameters  of 


FIG.  435. — Worthington  Motor  Driven  Ten-stage  Pump. 

impellers  could  be  used  to  get  the  necessary  velocity,  but  the 
friction  of  the  water  in  the  impeller  passages  and  on  the  im- 
peller back  would  be  very  great  and  the  width  of  passages 


FIG.  436. — Worthington 


Eight-stage  Turbine  Mine  Pump  with  450  H.P. 
Motor. 


at  discharge  would  be  very  small,  and  moreover  there  would  be 
erosion  of  the  vanes  by  the  water  under  high  velocity.  For 
these  reasons  about  150  feet  is  the  limit  for  each  stage  and  for 
great  heads  a  large  number  of  impellers  is  used.  Since  these  take 


578 


PUMPING  MACHINERY 


up  considerable  room  on  the  shaft  there  is  a  limit  to  the  number 
which  can  be  put  between  bearings  without  making  the  shaft 
very  large.  The  shaft  must  be  designed  with  regard  to  the 
critical  speed.  Ffve  or  six  stages  are  found  within  one  casing, 
but  when  more  are  required  two  casings  are  used  connected  in 
series. 

It  is  well  to  think  of  the  first  wheel  increasing  the  kinetic 
energy  and  the  potential  energy,  then  the  first  diffuser  changes 


FIG.  437. — Worthington  Fire  Boat  Turbine  Pump. 

the  increase  of  kinetic  energy  into  pressure,  so  that  when  the 
second  impeller  is  entered  the  water  has  the  same  kinetic  energy 
as  it  has  at  entrance  to  the  first  impeller,  but  with  a  great 
increase  of  pressure.  This  is  repeated  in  each  stage  and  at  the 
discharge  from  the  pump  at  the  last  stage  there  may  be  less 
kinetic  energy  than  at  entrance,  but  the  potential  energy  in  the 
form  of  pressure  is  very  great. 

When  heads  of  several  hundred  feet  are  required  for  fire 
service  the  pump  takes  the  form  shown  in  Fig.  437.     This  type 


CENTRIFUGAL  PUMPS  579 

of  unit  is  installed  on  the  New  York  and  Baltimore  fire  boats. 


FIG.  438. — Alberger  rooo-gal.  Underwriter  Fire  Pump,  Turbine  Type. 


FIG.  439. — Worthington  Four-stage  Boiler  Feed  Pump. 

These  pumps  are  also  used  as  fire  pumps  for  factories  and  the 
Board  of  Fire   Underwriters  has  issued  specifications  of    the 


580 


PUMPING   MACHINERY 


same  nature  as  those  given  in  Chapter  XI  for  reciprocating 
pumps  Covering  the  equipment  and  construction  of  centrifugal 
underwriter  fire  pumps.  Fig.  438  illustrates  a  looo-gallon  pump 
of  this  type.  The  fixtures  of  this  pump  are  quite  similar  to 
those  discussed  earlier  in  the  work. 

For  boiler  feeding  a  motor  driven  or  steam  turbine  driven 
stage  pump  may  be  used.     When   turbine   pumps   are  used 

there  is  an  absence  of  knocking  and 
shock  common  with  reciprocating 
pumps  because  of  the  steady  dis- 
charge, and  moreover,  as  will  be  seen 
later,  there  is  no  danger  of  excessive 
/Y\  pressure,  even  if  all  feed  valves  are 

HL  .  closed    on     the    boilers.      Fig.    439 

illustrates  a  four-stage  boiler  feed 
pump.  Pumps  similar  to  this  could 
be  used  for  house  or  elevator  service, 
being  controlled  by  a  switch  operated 
by  a  float  or  pressure  regulator. 

For  mine  sinking  or  for  use  where 
space  is  limited,  vertical  pumps  may 
be  installed,  using  electric'  motors  or 
steam  turbines  to  operate  them. 
These  pumps  have  enormous  capacity 
for  their  size,  and  operate  successfully. 
Fig.  440  illustrates  a  pump  built  to 
operate  under  1250  feet  head  in  a 
single  casing.  Should  the  water  to 
be  handled  be  acid  the  pump  is  built 
of  a  composition  to  resist  this. 

When  high-speed  pumps  are  built 
these  are  made  with  double  runners 

to  eliminate  end  thrust.  Fig.  441  shows  a  section  through 
one  of  these  a§  built  by  Worthington  and  shown  in  Fig. 
437.  In  this  pump  the  suction  water  for  one  side  is  carried 
through  the  open  spaces  between  the  channels  of  the  diffuser 
ring  of  Fig.  431  and  enters  a  well-rounded  cavity  leading  to 


FIG.  440. — Worthington  Turbine 
Sinking  Pump. 


FIG.  441. — Worthington 


>eed  Pump. 


(To  face  page  580) 


CENTRIFUGAL  PUMPS 


581 


FIG.  442. — Alberger  Two-stage  Volute  Pump,  Engine  Driven. 


FIG.  443. — Alberger  Standard  Volute  Pump,  Vertical  Shaft. 


582  PUMPING  MACHINERY 

the  impeller.  In  this  way  the  impeller  receives  water  on  each 
side.  The  water  from  the  diffuser  is  then  delivered  to  the  suc- 
tion of  the  second  stage  and  is  delivered  from  the  diffusion 
ring  of  this  stage  into  the  discharge  channel.  The  velocity 
head  in  this  channel  is  so  small  compared  with  the  pressure 
head  that  there  is  not  the  need  of  forming  the  usual  volute  to 
reduce  the  losses  and  so  the  discharge  is  taken  up  on  each  side 
of  the  diffuser  to  a  discharge  flange  in  the  center  in  a  concentric 
channel. 

Fig.  442  illustrates  a  two-stage  volute  pump  built  by  Alberger. 
In  the  figure  the  discharge  is  taken  from  the  casing  at  an  angle 
to  the  horizontal.  The  general  lines  of  the  pump,  the  feet  for 
attachment  to  bed  plate,  the  water  seal  for  the  suction  stuffing 
box  and  the  well-supported  bearings  are  to  be  noted.  The 
volute  pump  is  not  used  so  often  in  stages,  but  there  is  no  reason 
why  this  cannot  be  done,  as  is  shown  in  the  illustration.  The 
volute  pump  may  be  applied  with  the  shaft  in  a  vertical  posi- 
tion as  shown  in  Fig.  443.  Such  a  construction  may  be  neces- 
sary on  account  of  the  lack  of  room,  or  the  motor  may  be  placed 
at  considerable  distance  above  the  pump  on  account  of  flooding. 

Another  use  for  the  centrifugal  pump  in  the  last  few  years 
is  in  connection  with  the  water  removal  from  jet  condensers.  Fig. 
444  shows  a  steam-engine  driven  centrifugal  pump  attached  at 
the  lower  end  of  an  Alberger  condensing  head.  The  condensed 
steam  and  condensing  water  are  removed  and  discharged  into 
the  atmosphere  by  the  pump.  The  air  is  removed  from  the 
head  by  means  of  a  dry  air  pump,  similar  to  that  described  in 
Fig.  368. 

In  order  to  eliminate  the  end  thrust  with  turbine  pumps  of 
several  stages  Mr.  C.  W.  Larner  has  recently  patented  the  method 
shown  in  Fig.  445.  In  this  arrangement  the  water  from  the 
second  diffusion  ring  is  passed  between  the  channels  of  the  first 
channel  ring  on  its  passage  to  the  second  channel  ring.  In  the 
figure  there  are  four  stages.  The  method  used  here  is  some- 
what similar  to  that  shown  in  the  high-speed  pump,  Fig. 
441,  although  in  that  figure  each  impeller  is  balanced  by  being 
a  double-flow  impeller  and  half  of  the  water  for  the  second  stage 


CENTRIFUGAL   PUMPS 


583 


is  carried  through  the  channels  in  the  diff users.  In  Fig.  445, 
the  impellers  are  single-flow  impellers;  consequently  with  the 
same  openings  at  the  suction  side  the  pump  will  handle  less 
water  through  a  greater  head  than  that  for  the  pump  shown 
in  Fig.  441. 

The  method  used  by  Sulzer  to  accomplish  this  is  shown 


FIG.  444. — Alberger  Centrifugal  Condenser. 

in  Fig.  446.  This  has  been  used  for  many  years.  The  method 
is  to  take  the  discharge  from  the  first  stage  and  pass  it  through 
the  open  spaces  in  the  two  diffusers,  Fig.  431.  The  water  is 
then  brought  in  on  the  opposite  side  of  the  second  impeller 
from  that  ,used  on  the  first.  This  balances  the  pressure  due 
to  impact,  since  the  velocity  of  the  water  in  an  axial  direc- 


584 


PUMPING  MACHINERY 


tion  is  the  same  at  each  of  these  points  on  the  second  im- 


peller. 


Rateau  eliminates  the  end  thrust  for  one  condition  of  running 


FIG.  445. — Larner  Pump. 


FIG.  446. — Sulzer  Pump. 

by  cutting  off  one  side  of  the  impeller  as  shown  in  Fig.  447, 
and  allowing  the  unbalanced  force  on  the  side  remaining  to 
balance  the  pressure  at  entrance.  This  means  considerable 
leakage,  however,  and  moreover  the  wheel  for  a  part  of 


CENTRIFUGAL  PUMPS 


585 


one  side  acts  as  an  imencased  wheel,  which  means  more  fric- 
tion. 

In  the  three-stage  pump  of  Worthington,  Fig.  448,  the  Jaeger- 
Worthington  method  of  holding  the  impeller  in  a  central  posi- 
tion is  illustrated.  In  this  pump  the 
edge  of  the  diffuser  at  A  is  cut  on 
each  side  so  that  should  the  impeller 
move  to  the  right  the  increased  clear- 
ance space  allows  more  water  to  leak 
out  on  that  side  than  would  be  cared 
for  by  bushing  rings  at  B.  As  a  result 
the  water  "  backs  up  "  on  the  right 
side  and  the  pressure  from  this  forces 
the  impeller  to  the  left.  Motion  to  the 
left  would  result  in  an  excess  of  press- 
ure on  that  side.  To  permit  this  free 
action  the  thrust-bearing  grooves  are 
made  in  a  sleeve  which  has  a  slight 
play.  There  is  no  pressure  on  the 
thrust  bearing  until  the  sleeve  is 
brought  to  one  end  of  its  travel  or  the  other.  In  this  way  the 
thrust  is  carried  by  water  on  the  backs  of  the  impellers  and 
the  impeller  s*hould  take  such  a  position  that  the  excess  leakage 
on  one  side  would  care  for  the  thrust. 

In  the  pump  of  Weise  and  Monski  the  impellers  are  divided 
into  two  groups,  one  group  with  the  inlets  on  the  right  and  one 
set  with  the  inlets  on  the  left.  After  passing  through  the  first 
group  the  water  is  discharged  through  a  passage  in  the  casing 
leading  to  the  other  end  of  the  pump,  where  it  enters  the  inlet 
of  the  first  impeller  of  the  second  group.  This  is  the  equiva- 
lent of  having  two  pumps  with  their  shafts  connected,  one  pump 
discharging  into  the  other  through  an  outside  connection  and 
the  inlets  to  the  impellers  on  one  pump  turned  in  the  opposite 
direction  from  those  on  the  other.  Fig.  436  shows  such  an 
arrangement. 

In  the  Schwartzkopff  type  of  pump  and  in  the  later  Sulzer 
pumps  a  balancing  piston  is  used  at  the  end  of  the  pump  sub- 


FlG'  44'~ 


Centrifu^al 


586 


PUMPING   MACHINERY 


CENTRIFUGAL  PUMPS 


587 


FIG.  449. — Section  of  a  Buffalo  Balanced  Two-stage  Pump. 


FIG.  450.— Section  through  Buffalo  Pump. 


588 


PUMPING  MACHINERY 


ject  to  the  discharge  pressure  or  a  portion  of  it.     The  pressure 
on  the  piston  balances  the  thrust. 

Fig.  449  gives  a  section  through  a  two-stage  Buffalo  pump, 
showing  their  method  of  providing  for  end  thrust  while  the 
cross-section,  Fig.  450,  shows  the  form  of  passages  used  to  take 
the  water  to  the  second  impeller  inlet.  Fig.  451  shows  one  of 


FIG.  451. — Method  of  Assembling  Buffalo  Multi-stage  Pumps. 

these  pumps  partially  dismantled.  The  outer  part  of  the  dif- 
fuser  vanes  B,B  do  not  fit  against  the  shell,  but  are  bolted  by 
tie  bolts  G  to  the  return  chamber  D  and  the  diaphragm  C, 
which  reaches  to  the  floating  ring  around  a  bushing  on  the 
shaft.  The  return  plate  D  and  diaphragm  C  have  a  forced 
fit  into  the  casing.  The  bolts  /  are  jack  bolts  used  in  forcing 
out  the  diaphragm  C  and  return  plate  when  necessary  to  examine 
the  interior.  In  this  arrangement  of  pump  it  is  not  necessary 
to  disconnect  the  suction  or  discharge  pipe  when  repairing  the 
interior.  Passages  cast  in  the  casing  deliver  water  from  the 


CENTRIFUGAL  PUMPS 


589 


left-hand  diffuser  B  to  the  return  chamber  D  on  the  right.     These 
are  seen  in  Figs.  449  and  450. 

Fig.  452  shows  one  of  the   Buffalo   pumps  with  a  vertical 


FIG.  452. — Buffalo  Vertical  Underwriter  Fire  Pump. 

shaft  equipped  as  an  underwriter's  fire  pump.  The  pump 
would  be  driven  by  a  motor  through  the  coupling  shown  in 
Fig.  453 .  This  type  of  coupling  is  one  in  which  pins  on  one  flange 


590 


PUMPING    MACHINERY 


drive  the  other  flange  by  means  of  rubber  bushings.  Such  an 
arrangement  furnishes  a  yielding  medium,  so  that  there  is  some 
chance  for  self  alignment  when  the  shafting  is  deflected  in  part 
of  the  complete  machine.  In  most  installations  there  must 
be  a  flexible  coupling  between  the  motor  and  the  pump  on 
account  of  change  of  alignment  at  high  speeds. 

The  Allis-Chalmers  pumps  built  for  the  high-pressure  fire 
service  in  New  York  demand  mention  as  another  type  of  multi- 
stage pump.  These  pumps,  Fig.  454,  are  built  with  the  passages 
in  the  impellers  so  formed  that  the  outflow  is  in  an  axial  direc- 
tion. This  form  of  pumpi  is  known  as  the  Gelpcke-Kugel  form. 
These  pumps  on  test  gave  efficiencies  from  72  to  79  per  cent. 


FIG.  453. — Coupling. 

There  is  some  objection  to  the  discharge  as  shown,  in  that  the 
leakage  is  apt  to  be  excessive. 

Testing  Centrifugal  Pumps.  The  centrifugal  pump  is  often 
designed  to  run  at  a  fixed  speed,  being  driven  by  an  electric 
motor.  The  amount  of  power  used  and  the  efficiency  will  vary 
with  the  quantity  of  water  pumped,  and  for  that  reason  tests 
are  made  to  determine  the  characteristics  of  the  pump. 

The  characteristics  desired  are  the  curves  of  power,  of  pres- 
sure and  of  efficiency,  plotted  against  the  quantity  of  water 
pumped.  To  obtain  these,  measurements  are  made  as  shown 
in  the  diagram,  Fig.  455. 

The  quantity  of  water  may  be  determined  by  a  Venturi 
meter,  a  Pitot  tube,  a  calibrated  nozzle  or  a  weir.  The  head 
against  which  the  pump  is  working  is  determined  by  two 
gauges — a  suction  gauge,  and  a  discharge  pressure  gauge,  or 


CENTRIFUGAL  PUMPS 


591 


592 


PUMPING  MACHINERY 


CENTRIFUGAL  PUMPS  593 

the  suction  may  be  attached  to  one  end  of  a  mercury  U  tube 
while  the  discharge  is  connected  to  the  other  end.  The  power 
input  is  determined  by  knowing  the  power  and  efficiency  of 
the  electric  motor,  by  indicating  the  engine  used  to  drive  the 
pump  or  by  the  use  of  a  dynamometer.  The  arrangement  of 
the  apparatus  for  such  a  test  is  seen  in  Fig.  455,  in  which  there 
are  several  instruments  for  measuring  the  water.  The  suction 
gauge  at  A  is  usually  a  mercury  gauge.  The  pressure  gauge 
at  B  may  be  a  Bourdon  gauge.  The  Venturi  meter  at  C,  the 
Pi  tot  tube  at  Z),  the  nozzle  E  with  its  pressure  gauge  F,  or  the 
weir  G  may  be  used  to  determine  the  water  used.  The  Pitot 
tube  may  be  applied  simply  by  using  a  tube  flush  with  the  pipe 
surface  for  the  static  pressure  tube  and  sliding  the  velocity 
tube  across  the  pipe,  making  a  traverse.  With  a  sharp,  small 
opening  in  the  velocity  tube  and  the  tube  kept  parallel  to  the 
axis  of  the  pipe  the  constant  for  the  tube  is  unity  so  that 


h  is  the  difference  between  the  pressures  in  the  litot  tube  and  Lhe 
static  tube. 


q=  I     2irrdrV  2gk  =  S.o2X27r  I     rvhdr.    ;.     .     (75) 
Jo  Jo 

Hence  if  \/h  is  multiplied  by  r  and  the  product  is  plotted  on 
a  straight  base  for  different  values  of  r  from  o  to  ro  the  area  of 
this  curve  when  multiplied  by  16  04  ?r  will  be  the  quantity  in 
cubic  feet,  since  all  measurements  are  in  feet. 

The  formula  for  the  Venturi  meter  is 

'  Q~kVZ?=AJV*g(Hl-H*>>    '    "     '     (76) 

where  A1  and  A2  are  the  areas  of  the  sections  of  the  meter  at 
which  the  pressures  are  HI  and  H2  feet  of  water. 
The  formula  for  the  calibrated  nozzle  is 

Q  =  k\/h, 

where  h  is  the  head  shown  by  the  gauge  at  F  measured  in  feet. 
For  the  rectangular  weir  the  formula  of  Francis, 

.....     (77) 


594  PVMP1NG  MACHINERY 

may  be  used  with  fair  approximation  to  the  correct  value.     In 

v2 
this  H  is  the  head  on  crest  of  weir  in  feet  and  h=  —  is  the 

2g 

head  of  velocity  of  approach  and  b  the  breadth  of  the  weir.  All 
measurements  are  in  feet.  The  above  is  applicable  to  a  weir 
with  end  contractions.  If  contractions  are  suppressed  the  term 
o.2H  is  omitted. 

When  the  quantity  of  water  is  not  large  a  triangular  weir 
may  be  used,  with  success.  This  notch  is  made  with  sides  at 
right  angles  and  for  small  quantities  the  head  is  much  higher 
than  for  the  rectangular  weir,  and  for  that  reason  there  is  less 
error  in  measuring  the  acting  head.  The  formula  for  this  weir  is 

«?=o.3o5ff5/2.  ..  .  '.  ;.';.  ..   .    (78) 

The  constant  has  been  studied,  and  although  it  does  vary 
the  limits  are  not  far  apart,  being  0.291  at  0.7",  0.306  at  ij" 
nnd  0.303  at  5".  In  this  formula  H  is  head  on  crest  measured  in 
inches;  in  all  of  the  above  Q  is  measured  in  cubic  feet  per  second. 

The  head  acting  not  only  includes  the  suction  and  discharge 
head,  but  it  should  include  the  velocity  head  acting  at  the  time. 

Vo2 

Having  the  total  head  H  equal  to  hi  +h2  +  —  and  the  quantity 
of  water  Q,  the  useful  power  becomes 


In  this  w=  weight  of  i  cubic  foot  of  water,  which  is  deter- 
mined by  actually  weighing  a  known  quantity  of  water,  by 
the  use  of  a  hydrometer,  or  by  taking  the  temperature  and 
referring  to  a  table  of  weights  of  water. 

The  efficiency  of  the  pump  is  determined  when  the  applied 
power  is  found  by  means  of  a  dynamometer  or  calibrated 
electric  motor. 

HP 


The  results  of  a  test  are  usually  plotted  with  the  quantity 
of  discharge  as  abscissae  using    head,   efficiency,  and   applied 


CENTRIFUGAL   PUMPS 


595 


power. as  ordinates  of  three  separate  curves.  Fig.  456  shows 
such  curves  from  a  Worthington  8"  volute  pump.  In  this 
pump  the  head  remains  constant  for  a  considerable  time  as 
the  discharge  valve  is  opened,  so  that  more  water  is  discharged 
until  1800  gallons  per  minute  is  .reached,  when  the  head  begins 
to  drop  rapidly,  accompanied  by  a  decrease  of  efficiency  arid 
a  slight  increase  of  power.  It  is  to  be  noted  that  the  increase 
of  power  begins  to  fall  off  at  high  discharges,  due  to  the  decrease 


50 
.40 
30 
20 
10 
0 

w, 
a 

1 

'o 

G 

S 

^ 

*&* 

y 

n     • 

'HO 

ad 

*N 

N 

7U 

/ 

' 

"*x 

> 

\ 

\ 

\ 

V 

\ 

I 

<i 

•p. 

r^- 

&  —  • 

^—  — 

-^ 

p- 

1 

^ 

J^V 

^** 

\ 

o 

^ 

^^ 

1 

0     200    400    600   800   1000  1200  1400  1600  1800  2000  2200  2400  2600  2800  3000 
Capacity  in  Gallons  per  Minute. 

FIG.  456.— Test  Curves  from  an  8"  Worthington  Volute  Pump. 

of  head.  The  head  does  not  increase  as  the  discharge  valve  is 
closed  and  the  power  decreases,  due  to  the  decrease  in  quantity. 
The  efficiency  has  a  maximum  point  because  when  the  quantity 
changes  from  the  designed  quantity,  the  water  passes  through 
the  wheel  with  impact.  This  may  be  seen  when  it  is  remembered 
that  the  speed  of  the  pump  is  fixed  and  the  velocities  in  the 
channels  are  fixed  by  the  equation 


(81) 


596 


PUMPING   MACHINERY 


In  Fig.  457,  from  -a  Worthington  class  B  volute  pump,  the 
head  is  seen  to  be  changed  more  rapidly  than  in  the  preceding 
figure.  The  efficiency  reaches  a  higher  value  while  the  brake 
horse  power  of  the  motor,  which  is  delivered  to  the  pump, 
actually  decreases  as  the  quantity  passes  beyond  4000  gallons 
per  minute;  this  is  associated  with  a  head  of  44  feet.  The 
power  curve  is  quite  characteristic  of  a  centrifugal  pump,  as 


1000  2000  3000  4000  5000 

Capacity  P.M.  in  Gal's 

FIG.  457. — Test  Curves  of  a  10"  Worthington  Class  B  Volute  Pump. 

it  shows  how,  with  a  proper  design,  these  is  no  danger  of  over- 
loading the  motor  if  the  water  is  shut  off  (no  discharge)  or  if 
the  discharge  pipe  should  break  (maximum  discharge). 

Fig.  458  gives  the  characteristic  curve  from  a  six-stage 
turbine  pump  used  in  the  Brooklyn  High  Pressure  Station.  In 
this  figure  the  entire  range  from  zero  discharge  to  maximum 
discharge  under  no  head  is  shown.  The  horse  power  is  seen 
to  have  a  maximum  value  at  840  H.P.  and  the  maximum 
efficiency  is  seen  to  reach  .the  value  of  76  per  cent.  The 


CENTRIFUGAL  PUMPS 


597 


1000 


90090 


Head 


80080 


700 


600 


:o- 


500% 


40040 


300 
200 


100 


30— 


-20- 


\ 


<\ 


800  1600          2400  3200          4JJJ  •        4800          5600  6400 

Gallons  per  Min. 

FIG.  458. — Test  Curves  of  ic"  Six-stage  Worthington  Turbine  Pump. 


120 

.  —  - 

^—  - 

-^ 

, 

^^- 

^ 

100 
>£. 

X 

^\ 

^*~ 

^^ 

^—  * 

\ 

'3  80 

«S 

^ 

^^" 

\ 

W  70 
* 
g  _ 

v& 

&$ 

v^^ 

^-^*" 

*  —  •  — 

^v 

^ 

\ 

03  60 

^ 

^^ 

s* 

•^^ 

\s 

ss 

N  50 

a 

^ 

^ 

> 

X 

\\ 

-s40 

« 

^ 

V 

? 

\ 

\ 

w  30 

; 

? 

\ 

20 

/ 

r 

10 

/ 

0 

3 

2( 

X) 

4< 

30 

Q 

00 

8( 

K) 

10 

00 

n 

IX) 

1^ 

too 

11 

i(XJ 

180 

Gallons  per  Minute 
FIG.  459. — Test  Curves  of  an  Alberger  Single-stage* Pump. 


598  PUMPING   MACHINERY 

high  efficiencies  on  these  three  diagrams  are  important  to 
notice.  The  characteristic  curves  from  a  10"  single-stage  tur- 
bine pump  of  the  Alberger  design  is  seen  in  Fig.  459.  In  this 
the  complete  range  of  the  pump  is  given.  Fig.  460  gives  the 
excellent  results  of  a  pump  built  by  the  I.  P.  Morris  Co. 

These  curves  show  how  the  centrifugal  pump  is  suited  to 
conditions  under  which  the  head  would  vary  or  where  the 
quantity  would  vary.  In  the  case  of  emptying  drydocks 
where  the  head  changes  as  the  dock  is  emptied,  a  volute  pump 
would  be  used.  At  the  start  the  quantity  would  be  great  at 
a  very  small  head,  while  as  the  head  increases  the  quantity 
decreases,  and  in  some  cases  the  power  required  increases  up 
to  a  certain  point,  after  which  there  is  a  decrease  of  power  and 
quantity  as  the  head  increases. 

In  the  case  of  a  fire  pump  (Fig.  458)  the  variation  of  pressure 
with  the  change  in  quantity  is  evident;  with  4  4oo-gallon 
streams  the  head  is  900  feet,  while  with  8  such  streams  the  head 
is  reduced  to  705  feet.  When  12  4oo-gallon  streams  or  19  250- 
gallon  streams  are  used,  the  pressure  is  reduced  to  380  feet. 
In  none  of  these  cases  is  there  danger  of  overloading  the  proper- 
sized  motor.  In  selecting  a  pump  it  is  better  to  select  one  in 
which  the  characteristic  curve  of  head  gives  a  maximum  value 
at  zero  discharge,  as  there  will  be  no  difficulty  in  obtaining  the 
head  for  which  the  pump  was  designed.  If  it  is  desired  to  keep 
the  head  constant  at  the  various  discharges,  the  speed  of  the 
pump  will  have  to  be  increased,  and  with  this  the  power  of  the 
motor.  A  speed  characteristic  could  be  drawn  for  any  varia- 
tion of  head  from  the  head  curve  at  a  given  speed  by  remember- 
ing that  the  head  varies  as  the  square  of  the  speed. 

In  order  to  classify  pumps  a  similar  method  to  that  used 
in  turbine  classification  may  be  employed. 

Specific  Speed.  One  of  the  powerful  aids  in  the  classifica- 
tion of  turbines  for  the  purpose  of  design  so  that  a  given  design 
may  be  used  for  any  turbine  having  a  given  characteristic,  is 
the  specific  speed.  This  is  the  speed  of  a  turbine  of  design 
similar  to  a  given  turbine  but  of  different  scale,  such  that 
under  a  unit  head  it  would  develop  one  unit  of  power.  A 


CENTRIFUGAL   PUMPS 


599 


similar  characteristic  may  be  used  for  a  centrifugal  pump; 
here  the  specific  speed  of  a  pump  will  be  that  speed  at  which 


8     9     ft     a     a    9     £     3 


S     *     9     S    Jl     ft    9     ft     3 


a  pump  of  similar  design,  but  to  a  different  scale,  will  pump  one 
unit  volume  of  water,  through  a  total  head  of  unity.  To  derive 
the  formula  for  the  specific  speed  of  a  centrifugal  pump  a 


600  PUMPING  MACHINERY 

method  will  be  used  similar  to  that  used  by  Professor  L.  F. 
Moody  for  the  specific  speed  of  turbines  as  given  in  Zeitschrift 
f.  d.  gesampte  Turbinenwesen,  Sept.  10,  1909. 

Suppose  that  the  quantity  of  water  lifted  by  a  pump  run- 
ning at  Na  revolutions  is  Qa  and  the  total  head  is  kH.  If 
this  wheel  is  now  run  under  a  total  head  of  kiHi  the  speed 
will  be  changed  so  that 


because  Noc  Foe  VS. 

The  quantity  is  equal  to  the  area  multiplied  by  the  velocity, 
and  hence 


(83) 


If  now  the  scale  of  the  pump  be  changed  so  that  the  quantity 
of  water  is  changed  to  Q^  the  diameter  x  being  changed  from 
Da  to  ZY. 

Now 


hence 

' 


(84) 


When  Da  is  changed  to  ZV  the  speed  in  revolutions  per 
minute  will  be  changed  in  the  inverse  manner  because  the  veloc- 
ity remains  the  same  as  the  head  remains  constant,  hence 


but 


CENTRIFUGAL  PUMPS  601 

and 


Therefore 


If  AT"/  is  to  be  the  specific  speed,  the  total  head  K±Hi  will 
be  unity  and  the  quantity  Qi  will  be  unity. 
Hence 


(87) 


In  this  N  is  taken  as    R.P.M.;    Q,  cubic  feet  per  section  per 
second;  kH,  feet  head. 

This  could  be  derived  from  the  expression  for  the  specific 
speed  of  a  turbine  by  substituting  QH  for  HP  in  the  formula 
(for  turbines), 


N  -N 
iv«  —     ° 


This  formula  can  be  used  in  either  the  English  or  French 
system  of  units.  In  the  English  system  Q  is  in  cubic  feet  per 
second  and  H  in  feet,  while  in  the  French  system  Q  is  in  cubic 
meters  per  second  and  H  is  in  meters.  The  value  N8  in  the 
French  units  may  be  changed  to  N8  for  the  English  system  by 

multiplying  by  —  j. 

The  value  of  the  specific  speed  for  a  series  of  different  pumps 
has  been  plotted  in  Fig.  461  with  the  efficiency  of  the  pump  as 
ordinates,  and  from  it  the  specific  speed  for  the  best  efficiency 
may  be  found. 

The  specific  speed  of  a  pump  is  the  same  for  all  similar 
pumps,  and  for  a  given  quantity,  head  and  number  of  revolutions 
the  specific  speed  may  be  found.  This  immediately  tells  the 
designer  the  class  to  which  the  pump  belongs  and  what  assump- 
tions are  to  be  made.  Should  the  specific  speed  be  too  high, 
the  pump  could  be  made  of  several  single  stages  in  parallel, 


602  PUMPING  MACHINERY 

as  the  quantity  Q  in  the  formula  refers  to  the  quantity  for  one 
impeller  and  the  head  H  to  the  head  for  that  impeller.  For 
a  multistage  pump  the  quantity  Q  is  the  quantity  for  the  whole 
pump  while  H  is  that  for  one  stage.  Should  N8  in  a  desired 
plant  be  too  small  for  high  efficiency,  the  use  of  a  multistage 
pump  would  require  a  higher  specific  speed  and  so  give  a  better 
efficiency. 

DESIGNING  CENTRIFUGAL  PUMPS 

To  illustrate  the  methods  of  design,  applying  the  principles 
of  this  chapter,  a  pump  will  be  designed  to  lift  1000  gallons 
per  minute  through  350  feet  of  pipe  against  a  static  head  of 
300  feet,  using  a  motor  at  750  R.P.M.  or  1000  R.P.M.  It  is 
assumed  that  these  are  the  only  speeds  possible.  The  calcu- 
lations are  made  with  a  slide  rule. 
.  i.  Specific  Speed  and  Number  of  Stages. 

1000 
1000  gals,  per  nun.  =g—  —  —  77=2.23  cu.ft.  per  sec. 

From  (87), 


From  the  curve,  Fig.  461,  75  appears  to  be  a  good  specific 
speed,  hence 

IV«  -  /J  - 


#  =  36.6  ft. 

This  is  a  very  small  head  for  one  stage,  and  consequently 
a  smaller  specific  speed  might  be  used  or  a  higher  actual 
speed.  Decreasing  the  specific  speed  means  a  smaller  efficiency, 
but  one  which  is  not  reduced  very  much  if  proper  design  is  used, 
using  60  for  N8  and  the  higher  speed  of  1000  R.P.M. 

TT.     1000    / 

#*  =  -r—V2.23  =  25.0. 


CENTRIFUGAL  PUMPS 


603 


This  is  a  good  head  to  use,  requiring  3  or  4  stages.     Using 
3  stages  and  considering  the  losses  in  head  to  be  5%, 


0.90- 


0.80 


0.70 


0.60 


0.50 


30      40      50      60      70      80      90     100     110     120 


0.90 


0.80 


0.70 


0.50 


50      100     150     200     250     300     350 
R.P.M.-Speciflc  Speed 

FIG.  461.— Specific  Speed  Curves. 


604  PUMPING   MACHINERY 

This  is  not  as  good  a  specific  speed  as  could  be  obtained  if 
we  desired  to  use  another  stage.     In  that  case 


looox  2.23 


Could  the  speed  of  the  pump  be  changed  by  gears,  the 
actual  speed,  for  AT«=75,  and  3  stages  would  be, 


NHl    75  X 

N  =-7—  =-  —  +=~  =  1650. 


(fv 


This  is  the  speed  which  should  be  used  with  this  pump 
for  best  conditions  but  with  given  data  either  Ng=5&  or 
Na=46  will  be  taken.  To  keep  down  the  number  of  impellers 
the  second  of  these  will  be  employed.  From  the  curve  of 
Fig.  461,  the  efficiency  is  assumed  as  78%. 

2.  Size  of  Suction  and  Discharge  Pipe.    With  a  short  pipe 

/OQQ 

v8-=  8.02-J  —  -  approx.  from  (61) 
=  19.6    ft.  per  sec. 

sq-ins- 

<^=4ft  in. 

Use  5"  =4, 

A8  =  19.63, 

2.23X144 

v8  =  -  -  =  16.35  ft.  per  sec. 

19.63 

Loss  assuming  length  to  be  20  ft., 


fy  64.32 


CENTRIFUGAL  PUMPS  605 

If  discharge  pipe  is  of  the  same  size  the  loss  would  be 


This  loss  is  over  20%  of  net  lift  and  hence  a  larger  pipe 
will  be  taken  for  the  discharge,  say,  8  inches. 

^=50.26  sq.ins. 

6.38  ft.  per  sec., 


^,0.02x330(6.38)^ 

^zg         A       64.32 

This  is  a  permissible  loss. 
If  a  lo-inch  pipe  is  tried: 


The  saving  in  the  use  of  the  larger  pipe  is 

(6.2-2.1)  2.23  X62.5=io4HR 

If  the  pump  has  an  efficiency,  from  the  curve  at  a  specific 
speed  of  46,  of  80%,  if  the  motor  has  an  efficiency  of  90%, 
if  the  pump  runs  4000  hours  per  year  and  if  power  cost  I  cent 
per  K.W.  hour,  the  saving  per  year  is 


.oi  =$43.20. 


The  additional  cost  of  pipe  is 

[n(d  —  d}txl  1 

W-  -  Xcost  per  ton  [i  +  installing  factor], 

2OOO  J 

ro.287r(io-8)X33QXi2  .          ]r  . 

2000  -  $25.00]  [2]  =$174.60, 


606  PUMPING  MACHINERY 

Now  the  interest,  depreciation,  taxes,  and  insurance  on  this 
pipe  line  will  be  taken  at  9%,  so  that  the  yearly  charge  will  be 

0.09  X$i74.6o  =  $15.71. 

It  is  seen  that  the  saving  per  year  from  the  use  of  a  lo-inch 
pipe  over  an  8-inch  one  is  $43.20,  while  the  increased  cost 
is  $15.71  per  year.  The  lo-inch  pipe  will  be  used  and  con- 
nected to  the  volute  casing  by  a  reducer  from  8  inches.  A 
larger  pipe  should  be  tried  in  the  same  manner.  If  the  saving 
in  power  is  greater  than  the  increased  cost  on  investment, 
the  larger  pipe  should  be  used. 

To  aid  in  solving  problems  of  this  kind  the  table  on  page  607, 
taken  from  one  of  the  bulletins  of  H.  R.  Worthington,  is  given. 

In  this  table  the  amount  of  water  usually  carried  in  a  pipe 
of  a  certain  size  is  given  together  with  the  loss  in  head  in  1000 
feet  of  straight  pipe.  Since  the  drop  in  head  varies  with  the 
length  of  pipe  this  can  be  used  for  various  lengths.  For  large 
pipes  the  quantities  are  given  in  million  gallons  per  twenty- 
four  hours. 

v  2 
The  loss  in  the  wheel  is  assumed  to  be  0.3—  =  1.2. 

The  total  head  is  now 


=310.0  ft.  instead  of  315  assumed. 
3.  Probable  Brake  H.P.  of  Motor  and  Size  of  Shaft. 

'°  =  IO°  B.H.P. 


The  shaft  to  transmit  this  power  at  1000  will  have  a 
diameter  given  by  the  equation  below  if  no  bending  t>-  con- 
sidered: 


(32 
-V~ 


321,000  H.P. 


P.OOOXIOO 

\  1000  X  10,000 


CENTRIFUGAL  PUMPS 


607 


<3a<Da<Bac<cS<oC«a<3c!a)G<ua<ie<uC 
agaoaoaoaoaoaoao  a_o  a_o  a.g  a.o 

^2'oT^oJ2'y^2cjJ2or^ori2o^2"oJ2  o,23  uJ2  y.22  y 
ni'C  rt'C  rt'C  rt'C  rt'C  rt'C  3'C  rt'C  rt'C  rt'C  ST  «  C 


^2  o^2 
d  4J  rt  ' 


_ 

^H  fr  .  ^  fc  .  *  •"*  fr  . 

S     §     S 


o  o  t  o>o  o  t-  o  t  o  ** 


H  CM  fO   t--   "^00     O     t- 


t^          \O  tr>  M     HI  00 


T  C*OO  O    to     •    IO  t^-  to  M   10 
10  M  HI  O  O         V 


HI   ID  1000   O  CvCO   tvO   OOOOO^O 


N  O  N   'too  00  M  \n\o  >o  M 


oo  oo  1000  lONOOOOHO^OOO 
t^  N   l~-  TOO   C\    •  00     -CO     •   rr    •   O 


<5>oONO^OOOOOHiOOOOioO<NOtoOO\OO 
tONl  NMO\NOrO<O  tooO   O\00   «>  O  \O   M 


•  HI   er>  (f)in 


Mf>vOlH 


O     -to-vO     -d     -00     -tO-O     -tO-O     -O     •   to    • 
Mt        OO          ONHlt^HiNNt^TlOt^to  IOOO    Cl  00    N    O 


O     -lO-O     - 


T^8 


a3Ga3cq3ca3c«Ca3c43aa3ca3ca3Ga3ca)C 
a.o  a_o  a_o  a_o  a_o  a_g  aoaoaoao  a.o  a.g 

^2  w^2  o^2  oJ2  o^j  0^5  o^j  0^3  0^3  oJ5  o^j  o^  o 
3  C  d'C  rt'C  d'C  d'C  d'C  oJ'C  d'C  d'C  d'C  d'C  3'C 


to  >o\O  t  >ooO  t<5  O  O  O 

•     M        •     N        •     (O  TT       .     10       • 

00         ro        to  to        N 


fO         fO         fO 


608  PUMPING  MACHINERY 

On  account  of  bending  to  be  considered  later  as  well  as  the 
effect  of  critical  speed,  the  diameter  will  be  taken  as 


=  2  ns. 


4.  Angular  Arrangement  of   Vanes  and  Areas  at  Entrance 
and  Discharge.     From  Newmann's  formulae  (36),  (46), 


cot/?. 


The  area  of  outlet  measured  on  the  circumference  is 


where  n=  number  of  vanes; 

t' = thickness  of  vanes; 
b2=  breadth  of  opening. 

For  a  given  diameter,  the  breadth  62  is  given  by  the  formula, 

,  &  <?2 x 

2==~7 —          — Tf — \ —    =~7 —          — Tf — \ — 4~o' 

\      2     cos  a2/   '      \      2     cos  a2/     ® 

If  now  H  is  small,  62  may  be  of  proper  size  with  a  small 
value  of  x  and  a  large  value  of  cot  ft,  but  when  H  is  large  x 
should  be  as  large  as  possible  and  cot  ft  as  small  as  possible, 
so  as  to  make  62  of  proper  value.  Now  » 


w2=x\  gKH. 

From  this  it  is  seen  that  it  is  better  to  keep  x  as  small  as 
possible  consistent  with  other  demands  in  order  to  make  w? 
small,  because  the  friction  of  water  against  the  impeller  depends 
on  this  term. 


CENTRIFUGAL   PUMPS  609 

On  examining  Fig.  413  it  is  seen  in  the  first  place  that 
when  /?=9o°,  x  =  i  for  all  values  of  a2  and  also  that  with  values 
of  /?  near  90°  there  is  little  change  in  x.  Consequently  for 
high  heads  /?  should  be  made  as  near  90°  as  possible.  It  is 
true  that  with  a  negative  value  of  a2  x  is  larger  for  smaller 

values  of  /?,  but  cot  /?  increases  so  much  that  — —z  might  not 

be  large.  When  the  pump  has  a  low  head  opposing  .flow,  x 
may  be  smaller  and  /?  smaller,  giving  cot  /?  a  larger  value.  For 
high-pressure  stages  /?  may  be  made  85°,  while  with  low-pressure 
stages  /?  may  be  40  to  50°.  The  value  of  «2  is  usually  negative, 
30  to  60°  being  used.  By  using  positive  values  of  a2,  the 
speed  of  the  pump  w2  may  be  decreased,  giving  what  may  be 
called  bending  back.  If  «2  is  made  o°,  all  values  of  /?  give 
x  =  i,  hence  w2  =  \  gKH.  For  pumps  in  which  there  is  no 
diffuser  /?  is  made  almost  o°,  so  that  u  may  be  small,  giving  pd 
as  large  a  value  as  possible,  since  there  is  no  good  method  of 
changing  velocity  into  pressure  in  a  volute  pump.  This  will 
mean  that  a2  has  a  large  negative  value,  say,  —70°. 

Using  in  the  problem,  which  is  a  high-pressure  stage  problem, 
a2=  —  60°  and  /?  =  85°,  #  =  1.073,  KH  =  io^\.  feet,  the  following 
results : 


w2  =  I.073V  32.16  x  104  =62.0; 
N  =  iooo  per  min.  =16.7  per  sec.; 
62 


=  0.59  ft.  =7  ins. 


From  (47)  and  (48),  using  Fig.  414,  ;  =0.0038, 


v  tegKH    ^0.0038  x  64.32  X 104  _ 
ur=—     -        =  —  =4-75 


\/32.i6xio4    ,- 


610  PUMPING  MACHINERY 

As  a  check:  «  cos /?=«r=  53.9X0.08716  =4.7, 

h  Q 2.23x144 

=  ~T~          nt'    \  =    ~~T~ 

4.7(71X14 


assuming  six  vanes  &  inch  thick  at  point. 

To  find  the  data  for  entrance  certain  assumptions  must  be 

made.     —  varies  from  ij  to  2j.     cr  may  be  made  equal  or 

nearly  equal  to  «r.  It  is  better  to  have  ur<cr,  as  this  means 
that  the  velocity  at  discharge  is  small. 

Let  cr  =  i. 75 ur  =  1.75x4.7  =8.22. 

This  is  the  velocity  c  since  «=o°.  If  the  water  entering 
the  impeller  have  this  same  velocity  the  area  between  the 
hub  and  the  shrouding  or  wall  of  the  impeller  must  be  given 
by  the  equation, 

_^_<?_2.23Xi44_      Qs    ins 
f  h  c          g  22         39-°  sq.ms. 

5.  Diameters  of  Impeller.  The  hub  of  the  impeller  may  be 
taken  to  be  £  inch  thick,  giving  the  diameter, 


Ah=g.6  sq.ins. 
The  area  of  circle  at  inner  end  of  the  shrouding  will  be 


D.  =7.82=7!. 

Using  8  inches  for  the  diameter  at  the  edge  of  the  vane 
of  the  impeller  gives 

r2     14 

—  =-£-  =  1.75,  a  workable  ratio. 


CENTRIFUGAL  PUMPS  611 

Now  cr  =  i.75wr  =  i.75X4.7=8-22; 

W2          4   r 

•i-Ew-y6*-*";;    .  .  Jj    . 

^i  =8  ins.; 


cos  a  i/  0.23 


1.66  ins. 


6.  Alternate  Graphical  Method.     The  above  dimensions  may 
be  found  in  another  manner.     From  Eq.  (30), 


V        A  A 

The  ratios  —  ,  —  ^,  —  ^,  and  the  quantities  «i,  /?,  and 

^1      ^1       ^d 

are  assumed  from  practice.     Let 


and  ^  for  the  impeller  and  for  the  diffuser  each  0.15.     Then 


/ 64.32x100 

~\  2x1.75x1x0.996x0.966-0.3-^" 


In  Fig.  462  ABC  is  taken'  of  such  a  length  that  AB  repre- 


612 


PUMPING  MACHINERY 


sents  ri  and  AC,  r^==~ 


and  BD  is  drawn  perpendicular 

to  AB,  and  BE  is  drawn,  making  an  angle  of  75°  (&i  =75)  with 
AB.  Lay  off  BE  equal  to  37.7  feet  per  second.    A  perpendicular 

J 


FIG.  462.  —  Graphical  Design. 

ED  from  E  on  BD  will  fix  the  velocities 


£>£  =36.4  and 


A  line  is  drawn  from  A  through  D  cutting  a  perpendicular 
to  AC  from  C  in  the  point  F.  FC  is  equal  to  w2.  From  the 
figure  this  equals  63.7,  FG  is  laid  off  at  85°  to  a  line  parallel 
to  AC  and  is  made  equal  to  69  feet  per  second.  This  gives 
CG=c2=6.2,  «2=  -33°- 


CENTRIFUGAL   PUMPS  613 

From  the  figure  ^,=5.26,  cr=c=g.8; 
63.7x12 


2.23X144 
b2  = p 7— TH  =1.40  ms.; 


2.23X144 


7.  Method  for  Volute  Pumps.  In  the  design  the  problem 
has  been  to  find  the  dimensions  of  a  turbine  pump,  one  in 
which  a  diffuser  lias  been  used  in  which  u  is  reduced  to  u3 
by  the  enlarging  channel.  As  a  result  of  this,  u  has  been 
made  quite  large  and  c2  has  been  small.  This  is  due  to  the 
assumption  of  £  =  85°,  or  some  large  value,  and  a2=—  60°. 
If  now  there  is  not  sufficient  space  to  reduce  u  to  u3,  or  if  the 
diffuser  is  to  be  eliminated,  then  u  is  made  small  by  selecting 
a  large  negative  value  for  a2  and  a  small  value  for  /?,  as  shown 
in  Fig.  463.  Such  a  pump  is  usually  employed  when  the  head 
is  low.  In  this  case  the  velocity  u  from  the  impeller  should  be 
radial  and  it  is  to  be  changed  in  the  whirlpool  chamber  to  u% 
by  gradually  enlarging  the  width  and  by  the  increase  in  the 
circumference.  When  the  volute  is  reached  the  velocity  is 
vr  in  a  tangential  direction.  The  loss  then  is 


8.  Form  of  Impeller  and  of  Vane  Curves.  Using  the  results 
of  the  latter  method  the  following  data  will  apply  to  the 
pump: 


614 


PUMPING  MACHINERY 


FIG.  463. — Diagram  for  Volute  Pump  and 
Low-head  Turbine  Pump. 


FIG.  464.— Section  of 
Impeller. 


202=63.8  ft.  per  sec. 
w\  =36.4  ft.  per  sec. 
c=cr=g.8  ft.  per  sec. 
Ci=37.8  ft.  per  sec. 
bi  =1.27  ins. 

C2=6.2. 

^=60.3. 


ns. 
ins. 


CENTRIFUGAL  PUMPS 


615 


The  data  will  be  first  applied  in  laying  out  an  impeller 
which  has  pure  radial  action,  such  as  used  in  Figs.  423,  429,  430. 
The  cross-section  will  be  laid  out  as  shown  in  Fig.  464.  The 
shaft  diameter  will  be  investigated  after  the  weight  of  the 
impeller  has  been  determined.  The  next  important  step  is 
to  lay  out  the  vanes  of  the  impeller  and  diffuser.  There  are 
several  methods  which  may  be  employed. 

In  Fig.  465  the  vane  is  drawn  as  a  parabola.     At  A  the 


FIG.  465. — Parabolic  Vanes. 

angle  «i  is  laid  off  from  the  radial  direction  and  then  the 
point  B  is  so  chosen  that  the  line  making  the  angle  a2  with 
the  radius  will  intersect  the  line  from  A  at  C,  about  a  mean 
radius  between  r±  and  r2.  These  lines  are  tangents  to  the  vane 
curve  at  entrance  and  exit  and  to  put  in  the  curve  between 
the  points,  the  lines  AC  and  CB  are  divided  into  the  same 
number  of  parts  and  lines  are  drawn  connecting  similar  points, 
the  top  point  of  AC  being  connected  to  the  top  of  CB,  etc. 


616 


PUMPING   MACHINERY 


These  lines  axe  tangents  to  the  parabola  and  hence  their 
envelope  will  be  the  desired  curve.  When  DE  is  drawn  at 
one-sixth  the  circumference  from  BA  it  is  seen  that  there  is 
considerable  change  in  angle  from  A  to  £,  and  hence  an  extra 
vane  GF  is  put  in.  Since  these  vanes  converge  so  rapidly  as 
seen  at  H  there  is  some  danger  of  interference  with  the  flow 
of  water.  Consequently  these  vanes  are  usually  so  drawn  that 


FIG.  466. — Involute  Curves. 

there  is  a  portion  of  the  vane  just  opposite  to  the  corner  of 
the  next  vane  which  is  parallel  to  it. 

Such  a  result  may  be  obtained  by  using  the  involute  as 
the  curve  at  entrance  and  exit  when  possible.  At  A,  Fig.  466, 
the  angle  a\  is  laid  off  as  before  and  a  perpendicular  AB  to  this 
line  is  drawn.  The  circle  tangent  to  this  perpendicular  will 
be  the  base  circle  of  the  involute  AC.  Another  involute  drawn 
from  D  will  be  parallel  to  the  first  involute,  because  the  curves 
at  the  points  C  and  D  have  the  same  centers  of  curvature. 


CENTRIFUGAL  PUMPS 


617 


If  the  involutes  were  carried  further  they  would  be  parallel 
and  at  the  constant  distance  CD  apart.  At  the  point  C  the 
water  is  moving  parallel  to  that  entering  at  D  and  consequently 
there  is  no  tendency  for  the  water  at  the  entering  corner  to 
be  affected  by  the*  interference  of  the  previous  vane.  At  F 
and  H  the  water  is  so  moving  that  it  is  parallel  to  water  entering 
the  impeller  directly  below  it  at  /  and  /,  when  the  water. at 


FIG.  467. — Involute  Curves. 

/  and  /  is  entering  at  an  angle  ai  to  the  radius  at  the  point 
in  question. 

When  the  same  method  is  used  at  outflow,  it  is  found  that 
the  circle  tangent  to  the  perpendicular  MN,  perpendicular  to 
the  line  at  an  angle  of  «2  with  the  radius,  has  its  point  of 
tangency  N  within  the  space  left  for  one  vane.  This  would 
mean  that  there  would  be  nothing  gained  by  the  use  of  the 
involute.  Intermediate  vanes  0  and  P  are  sometimes  intro- 
duced to  keep  the  outflow  in  the  proper  direction  and  by  placing 
0  so  that  it  falls  between  N  and  M  the  action  of  outflow  is 
similar  to  that  at  inflow.  These  vanes  are  only  carried  in  a 


618 


PUMPING  MACHINERY 


portion  of  the  distance  toward  inlet  as  it  is  not  desired  to 
obstruct  the  inflow.  The  involutes  at  inflow  and  outflow  are 
connected  by  a  curve  as  at  M ,  or  a  tangent  as  at  C. 

When  «2  has  a  larger  negative  value  the  involute  at  out- 
flow takes  the  form  shown  in  Fig.  467.  At  A  the  two  involutes 
are  so  far  apart  that  a  reverse  curve  has  to  be  used  in  joining 
them,  while  at  B  a  curve  of  the  same  form  and  curvature  could 
be  used.  At  C  the  involute  at  outlet  is  only  used  at  the  tip 


FIG.  468. — Circular  Arc  Vanes. 

while  a  tangent  is  used  between  the  parts.    The  form  at  B  is 
quite  common. 

In  Fig.  468  positive  values  of  «2  are  shown  as  well  as  the 
method  of  using  circular  arcs.  If  perpendiculars  to  the  lines 
of  flow  (at  angle  a\  to  radius)  are  erected  at  A  and  B  these 
intersect  at  C.  The  point  C  may  be  used  as  a  center  for  the 
arc  from  B  to  D.  This  curve  at  D  is  parallel  to  the  curve  A 
at  the  point  A  and  hence  there  is  no  tendency  to  interfere 
with  the  flow.  The  perpendiculars  at  E  and  F  at  outflow 
intersect  at  G,  which  is  the  center  for  the  arc  at  outflow.  The 
centers  at  outflow  and  inflow  lie  on  circles  from  the  center 


CENTRIFUGAL  PUMPS  619 

of  the  wheel  and  these  are  dotted  in.  After  finding  the  centers 
C  and  G  and  the  radii  CA  and  GE,  these  other  curves  are 
quickly  drawn  in  by  using  the  dotted  circles  on  which  the 
centers  must  lie. 

One  set  of  vanes  shows  how  the  circular  arcs  are  joined  by 
tangents  while  another  shows  the  use  of  a  curve.  At  K  a 
parabola  is  constructed  showing  the  method  of  using  that 
curve. 

If  the  line  HI  is  drawn  in  the  direction  of  flow  from'  /  it 
appears  as  if  there  might  be  a  dead  space  EHI  where  no  water 
would  flow,  but  this  would  be  filled  with  eddies  causing  loss. 
If  the  change  in  form  is  too  sudden  this  may  happen,  but  it  is 
to  be  remembered  that  this  water  is  under  pressure  and  it 
would  consequently  gradually  enlarge  if  the  curvature  is  not 
too  great. 

The  curves  drawn  in  the  preceding  figures  are  the  center 
lines  of  the  vanes  and  the  half  thickness  is  added  to  each  side 
of  the  center  line,  the  end  being  drawn  to  a  sharp  point. 

MIXED  FLOW  PUMPS 

The  figures  shown  are  applicable  to  pumps  in  which  there 
is  a  pure  radial  action.  If  it  is  expedient  to  have  some  axial 
action  at  the  center  on  account  of  a  desire  to  keep  the  outer 
diameter  small,  the.  vane  is  carried  into  the  center,  as  shown 
in  Fig.  469.  In  this  pump  the  peripheral  speed  at  inlet  changes 
for  the  various  parts  of  the  vane.  At  a  the  speed  is  w\a  at  b, 
Wib\  at  c,  wic,  etc.  The  velocity  c  should  be  the  same  at  all 
points  and  in  a  direction  normal  to  the  various  peripheral 
velocities.  To  have  this  result  a\  must  vary  over  the  inlet 
edge  of  the  vane. 

Before  carrying  out  the  steps  used  in  design  it  will  be  well 
to  consider  the  action  at  entrance.  Suppose  the  number  of 
vanes  is  assumed  and  the  thickness  of  metal  used  for  them. 
If  in  Fig.  469  the  inlet  edge  be  divided  into  four  parts  by  the 
points  i,  2,  3,  4,  5,  and  the  middle  points  a,  b,  c,  and  d  be 
marked,  the  lines  of  flow  aa" ',  bb" ',  cc" ,  dd"  may  be  approx- 
imated. These  stream  lines  or  lines  of  flow  may  be  considered 


620 


PUMPING  MACHINERY 


d"c"b"a" 


FIG.  469. — Section  of  Mixed  Flow  Impeller 


FIG.  470. — Velocity  Diagram  for  Various  Inflow  Points 


CEN  TRIP  VGA  L  PUMPS 


621 


to  be  on  conical  surfaces  at  the  points  of  entrance,  and  if  tangents 
are  drawn  from  the  points  a,  b,  c,  d  to  the  axis  of  rotation, 
these  lines  are  the  elements  of  the  various  conical  surfaces. 

The  peripheral  speed  at  the  points  a,  b,  c,  d  may  be  found 
as  shown  in  Fig.  470  where  oa,  ob,  oc,  etc.,  represent  the  radial 
distances  to  the  various  points.  wia,  wlb,  wic,  etc.,  will  repre- 
sent the  various  peripheral  velocities.  These  are  known  from 

w2,  since  Wix=w2 — -. 
From  Eqs.  (3)  and  (4), 


u< 


C22 


c2 


_±__L±?_Z^L2_L£i!  '     W__r        ~_ 

2g         2g        2g         --        <»«—--•'«*-_. 


FIG.  471. — Angular  Relations  on  Developed  Conical  Surfaces. 

Now  in  the  pump  considered  pd  —pb  is  the  pressure  difference 
between  that  in  the  discharge  space  and  that  in  the  suction 
space.  This  is  constant  for  all  of  the  suction  chamber  and 
hence  it  is  constant  for  all  points  of  entrance.  w2  and  c2  are 
constants  for  all  divisions  of  the  pump  and  the  losses  may  be 
considered  as  constant.  Hence 

Ci2  —  Wi  r2  =  constant. 

If  c  is  to  act  at  right  angles  to  wlx  the  expression  above 
is  equal  to  c.  c  is  therefore  a  constant  and  the  various  values 
of  c\  and  «i  at  a,  b,  c  and  d  are  found  as  shown  in  Fig.  470 
by  making  the  vertical  distances  all  equal  to  c. 

Fig.  471  is  now  constructed  by  developing  the  various 
conical  surfaces  and  drawing  on  these  developments  the 
various  vanes.  The  construction  for  point  c  is  given  as  illus- 


622 


PUMPING  MACHINERY 


trating  the  method  for  all  of  the  points.  From  the  center  o, 
Fig.  471,  oc  is  laid  off  equal  to  cc'  of  Fig.  469,  and  with  this 
as  a  center  the  arc  cc'"  is  drawn.  The  arc  cciv  is  now  drawn 
of  radius  ccv  from  Fig.  469  and  civ  is  determined  by  the  number 
of  vanes  used  in  the  circumference  (twelve  assumed).  If 
this  curve  is  then  rectified  and  placed  on  the  arc  of  the  conical 
surface,  the  pitch  distance  at  the  point  c  (Fig.  471)  is  found. 
The  angles  «ic  from  Fig.  470  are  now  laid  off  from  the  radial 
lines  at  c  and  c"'  and  then  the  involute  ccvl  is  drawn  in  and 
from  this  the  perpendicular  distance  between  the  vanes  cf"cvl 

is  found.  This  distance  less  the 
thickness  of  the  vanes  is  the  net  depth 
of  the  passage  or  c'"cvl  —t=dc. 

This  is  done  at  each  point  and 
then  the  lengths  of  the  outflow  edge 
are  rectified  as  in  Fig.  472,  giving 
the  line  a,  b,  c,  d,  as  the  length  of 
the  outflow  edge.  If  now  the  depths 
found  at  the  various  points  be  laid 
off  perpendicular  to  the  above  line 
at  their  respective  points  a  figure 
aa'd'd  is  found  which  represents 
practically  the  outflow  area.  This 
area  is  not  quite  equal  to  the  outflow 
area  as  this  is  a  warped  surface  which 
cannot  be  developed.  Now  the  quan- 
tity of  water  flowing  through  any 
element  of  length  is 


FIG.  472. — Curves  of  Area  and 
Quantity. 


or 


This  means  that  Q  is  represented  by  an  area  in  which  the 
ordinates  are  Cidc  and  the  abscissae  are  /.  Hence  if  the  depths 
d  are  multiplied  by  the  corresponding  velocities  GI  from  Fig. 
470  and  the  product  is  laid  off  at  the  various  points  of  the  edge 


CENTRIFUGAL  PUMPS  623 

in  Fig.  472,  the  figure  add" a"  is  found,  the  area  of  wnich  repre- 
sents the  quantity  of  water  flowing. 

Now  in  the  formula  (when  inflow  is  at  o°), 

Work  =  uw2  sin  /?. 

The  work  per  pound  of  water  is  the  same  at  all  parts  of  the 
impeller.     Hence  when  this  is  put  into  the  form 

Work  =Qwu2 — -.—  sin  /?  sin  a\, 

YI  A  i 

where  ri,  ^4i,  and  «i  change  for  different  parts  of  the  impeller, 
the  quantities  r\,  A\  and  a\  should  refer  to  a  point  at  the  cen- 
ter of  gravity  of  the  outflow  area.  To  show  this  suppose  that 
the  area  of  discharge  is  divided  into  a  series  of  elements  JAex, 
such  that  the  quantity  discharged  is  the  same  in  each.  Then 
=c\xAA\x=K  and 

Work  =  2wcixAA\x  u2—  —.—  sin  /?  sin  aix. 

r\ 


Now  Ci  approximately  varies  as  r,  as  may  be  seen  from 
Figs.  470  and  472  and  sin  ai  is  practically  constant. 

K"    ^KI^- 


Now  ZrAAlx  is  the  static  moment  of  the  inflow  area  and 
for  that  reason  it  equals  rc.gAi,  or  the  work  is  that  required 
by  the  water  if  it  all  entered  at  the  radius  of  the  center  of 
gravity  of  the  actual  area. 

To  find  the  center  of  gravity  of  aa'd'd  lay  off  the  curve 
dlvclvblvalv,  found  by  multiplying  each  ordinate  aa',  bb',  etc., 
by  the  distance  from  the  left-hand  end  of  the  figure  and  using 
these  as  ordinates  for  the  new  curve, 

Xae, 


The  area  of  this  curve  is  the  static  moment  of  the  original 


524.  PUMPING  MACHINERY 

area  about  the  left  corner.  This  area  divided  by  the  area  of 
the  original  curve  will  give  the  distance  of  the  center  of  gravity 
from  the  left  corner, 


edlvclvblvalv 
C'g'=      dd'a'a  ~ 


J 


fdc 


xdl 


In  designing  a  mixed  flow  impeller  the  method  of  procedure 
is  as  follows:  Assume  for  a  given  design  the  radius  at  the 
center  of  gravity,  as  rQ,  Fig.  469,  and  with  u2,  w2,  c,  and  aiy 
known  for  the  point  r0,  the  constructions  of  Figs.  462  or  463 
will  give  the  distance  dQ  and  the  velocity  cio.  The  length  of 
outflow  edge  is  now  found  approximately  by  the  formula, 

l=_Q_ 


This  is  then  used  in  Fig.  469  and  a  curve  drawn  equal  to 
this  in  length.  After  the  curve  is  assumed  the  various  points 
are  used  as  shown  in  Figs.  469,  470,  471,  and  472,  and  the 
actual  area  and  center  of  gravity  is  found.  If  now  the  center 
of  gravity  is  slightly  different  from  that  assumed,  or  if  the  area 

of  the  quantity  curve  is  greater  than  —  ,  the  length  or  shape 

of  the  outflow  edge  is  altered  to  give  the  correct  values. 
Several  trials  will  give  the  desired  e.g.  and  quantity. 

Since  there  is  a  slight  change  in  the  angle  <*i  and  moreover 
since  the  paths  are  of  different  projected  lengths  the  surface  of 
the  vane  is  rather  complex.  The  inlet  edge  is  often  contained 
in  a  radial  plane  so  that  the  curve,  Fig.  469,  is  seen  in  its  true 
length.  At  other  times  the  curve  is  in  a  plane  which  does  not 
pass  through  the  axis  and  sometimes  the  inlet  edge  is  a  non- 
planar  curve.  In  the  second  case  the  projection  of  the  curve 
on  a  plane  perpendicular  to  the  axis  would  give  a  straight 
line  passing  one  side  of  the  axis  while  in  the  last  case  the 
projection  would  be  a  curve. 

Since  the  curves  of  the  vanes  at  inlet  have  angles  ai  differ- 
ing slightly  from  each  other,  and  since  the  paths  to  the  outlet 


CENTRIFUGAL  PUMPS  625 

are  of  varying  length  the  outlet  line  may  not  be  parallel  to 
the  elements  of  the  outlet  cylinder,  although  by  properly  select- 
ing the  shape  of  the  paths  this  may  be  accomplished. 

Fig.  473  shows  a  first  set  of  curves  for  a  wheel.  Fig.  469 
and  the  cross-sectioned  part  of  Fig.  473  are  revolved  pro- 
jections of  the  vanes  as  if  the  complete  vane  was  in  a  radial 
plane.  The  stream  lines  are  not  in  the  position  shown.  To 
obtain  their  correct  position  the  following  method  is  used: 
If  the  quantity  curve  of  Fig.  472  be  divided  into  equal  parts, 
say  four,  by  the  lines  A,  B,  and  C,  these  lines  determine  the 
positions  along  the  outflow  edge  at  which  divisions  could  be 
placed  in  the  passage,  so  that  each  part  would  carry  one-fourth 
of  the  quantity  passing  through  the  impeller.  At  the  outlet 
edge  the  divisions  will  be  spaced  equally  because  c2  and  «2  are 
the  same  at  all  points.  If  now  the  constructions  of  Figs.  470 
and  471  be  made  for  each  of  the  points  a,  b,  c,  d,  and  e  of  Fig. 
473,  the  true  shape  in  projection  of  these  lines  may  be  found. 
The  points  at  the  half  pitch  across  the  bucket  have  been  used 
to  get  better  results.  The  points  will  all  lie  on  a  radius  15° 
from  OR  in  the  end  view  of  the  figure.  Consider  the  part  marked 
C  of  Fig.  471.  The  point  of  the  vane  on  a  radial  plane  15° 
from  the  entering  edge  of  the  involute  vane  will  be  at  a  dis- 
tance ab  from  involute  to  circle  at  the  edge.  This  point  is  at 
a  distance  ab=ccf  from  point  c  on  the  element  of  the  cone, 
but  when  this  cone  is  wrapped  into  position  the  true  position 
of  the  point  will  be  on  an  element  15°  from  the  original 
point  in  the  end  view.  Hence  if  c'  is  projected  over  to  c" 
in  the  end  view  and  this  is  swung  to  the  15°  radius,  the 
point  c'"  is  determined.  This  is  carried  over  to  the  side 
view  and  clv  is  then  determined  by  the  intersection  of  this 
projecting  line  and  a  perpendicular  to  the  axis  from  c' '.  In 
the  same  manner  cc^  is  made  equal  to  c'"cvl  and  Cilv  is  found 
from  the  point  c\"  in  the  end  view,  which  is  determined  by 
swinging  c\"  over  to  the  30°  radius.  The  intersection  of  the 
two  projecting  lines  gives  c\lv.  The  same  operation  is  used  for 
the  other  points.  From  the  first  two  points  of  these  various 
lines  in  the  end  view,  the  remainders  of  these  lines  are  sketched 


62G 


P  UMPIXG  MA  CH1NER  Y 


CENTRIFUGAL   PUMPS  627 

in  for  this  view  and  these  are  ail  brought  to  the  same  element 
of  the  discharging  cylinder  and  «2  is  made  the  same  for  each. 

To  construct  the  true  shape  of  the  various  stream  lines  in 
the  first  view  from  the  curves  just  sketched  in,  the  following 
method  is  used:  such  a  point  as  c4  is  revolved  to  c5,  carried 
over  until  it  strikes  the  stream  line  from  c  at  c&,  this  is  pro- 
jected down  until  it  cuts  the  projection  line  from  c4,  giving 
the  true  point  C?.  This  is  done  for  all  stream  lines,  and  if  the 
shape  of  any  is  too  complex  or.  has  too  much  reverse  curvature, 
a  new  set  of  lines  is  assumed.  After  these  two  projections  are 
made,  sections  of  the  vane  form  should  be  made  at  various 
angles  for  the  purpose  of  seeing  that  the  vane  will  have_-the 
proper  shape  in  all  directions.  Such  an  operation,  similar  to 
fairing  ship  curves,  is  important.  The  curves  seen  at  the  right 
of  Fig.  473  are  the  radial  plane  intersections.  Those  shown  as 
heavy  dotted  lines  in  the  end  view  are  those  on  planes  I,  II, 
III,  IV,  V,  and  VI  perpendicular  to  the  axis.  These  latter 
curves  are  those  used  in  constructing  the  core  boxes  from 
which  the  moulds  for  casting  are  made.  In  the  above  there 
has  been  no  thickness  of  vanes  shown.  After  finding  the 
curves  for  one  side  of  the  vane  a  similar  operation  could  be  used 
for  the  other  face. 

AREA  CURVE  THROUGH  BUCKET 

After  the  vanes  are  determined  and  drawn  in,  it  is  well 
to  find  the  area  at  various  points  along  the  middle  line  of  the 
bucket  (center  of  gravity  of  various  areas)  and  if  the  lengths 
between  various  centers  be  laid  off  as  a  base  line,  Fig.  474, 
and  the  areas  at  these  points  are  used  as  ordinates,  the  area  of 
this  figure,  by  the  theorem  of  Pappus,  is  the  volume  of  the  vane 
bucket. 

The  curve  so  constructed  should  gradually  increase  or 
decrease.  A  curve  with  considerable  change,  as  shown  by  the 
dotted  line,  is  objectionable,  as  this  means  frequent  changes 

in  velocity \vx=~T~)  and  changes  in  velocity  are  usually  accom- 


P  UMPING   MA  CHINER  Y 


panied  by  loss.  If  such  a  curve  as  the  dotted  one  is  found, 
the  axial  width  should  be  changed  to  bring  it  into  the  condi- 
tions shown  by  the  solid  line  or  the  vanes  are  thickened  up 


FIG.  474. — Curve  of  Passage  Area. 

by  the  use  of  back  vanes,  as  shown  in  Fig.  475.  These  are 
necessary  in  the  figure  shown  to  keep  the  velocity  more  nearly 
constant,  as  the  distance  between  the  vanes  is  too  great  at  the 
center. 


FIG.  475. — Back  Vanes. 


ABSOLUTE  PATH  OF  WATER 

The  area  of  the  curve,  Fig.  474,  between  the  inlet  point  and 
any  other  point  is  the  volume  of  the  bucket  to  that.  point  and 
since  AiCi  is  the  quantity  of  water  entering  the  bucket  per 
second,  this  volume  divided  by  A^c^  is  the  time  taken  for  a 
particle  of  water  to  move  from  the  entering  edge  to  the  point 
considered, 


*~ 


vol. 


x     area  x 


CENTRIFUGAL  PUMPS 


629 


To  find  the  absolute  path  of  a  particle  of  water  passing  over 
the  vane  av  of  the  Fig.  476,  the  times  taken  to  pass  to  the  points 
v,  x,  y,  z,  and  a  are  found,  as  shown,  from  a  figure  similar  to 
Fig.  474.  The  points  vy  x>  etc.,  are  then  carried  in  radially 
to  v',  x',  y',  etc.,  and  from  these  points  the  distances  v'vi, 


FIG.  476. — Absolute  Path  of  Water. 


#iV,  yfyi'i  -  •  •  j  equal  respectively  to  w^,  w\tx,  w\ty,  etc., 
are  laid  off.  These  points  determine  the  positions  of  the  radii 
when  the  water  reaches  the  point  in  question.  Hence,  if  from 
x  an  arc  is  drawn  intersecting  the  radial  line  from  x\  this 
determines  the  point  Xi  of  the  absolute  path.  In  this  manner 
#i>  ^i>  yi>  #ij  and  Vi  are  found,  giving  the  absolute  path  of  the 
water  through  the  impeller.  This  path  should  be  a  smooth 


630  PUMPING  MACHINERY 

curve  without  sudden  changes  'of  curvature.  The  tangent  to 
it  represents  the  absolute  direction  of  the  water  at  any  instant, 
hence  at  outflow  the  angle  formed  with  the  radius  is  /?,  while 
at  entrance  the  angle  is  a. 

9.  Diffusion  Chamber.  Having  the  impeller  designed  the 
next  step  is  to  design  the  diffusion  chamber.  The  angles  of  the 
vanes  of  the  diffuser  are  /?  at  entrance  and  ad  at  discharge. 
Since  the  water  is  to  discharge  into  the  volute  cas'ng  and  travel 
in  that  chamber  in  a  tangential  direction  it  is  advisable  to 
make  ad  as  large  as  possible.  The  velocity  of  discharge  v^ 
is  given  by  the  equation, 

O 


-^' 

cos  ad  \ 


cos  «* 


When  ad  is  made  large  Dd  must  be  large  and  td  must  be 
made  very  large  if  vd  is  made  small.  This  cannot  in  general 
be  done,  and  in  many  cases  ad  is  made  small,  being  o°  in  some 
cases.  This  gives 

Q 


Dd  may  be  so  large  in  this  case  that  if  td  is  increased 
over  the  t  at  exit  from  the  impeller,  the  net  area,  even  after 
the  enlargement  of  tr  to  give  a  better  flow,  to  care  for  the 
supporting  bolts  or  to  make  a  passage  as  is  done  in  the  Worth- 
ington  pump,  is  so  large  that  vd  is  quite  small. 

In  Fig.  477  six  vanes  of  the  diffuser  have  been  drawn  as  cir- 
cular arcs  with  ad  and  /?  as  the  angles  at  the  two  ends.  On 
one  of  the  center  lines  a  plain  vane  has  been  drawn  with  its 
sharpened  end,  on  a  second  one  the  vane  has  been  constructed 
to  make  a  channel  for  a  supply  to  the  next  stage  for  balancing, 
while  the  third  shows  the  enlargement  of  the  vane  to  make  a 
more  gradual  change  in  area  and  to  form  a  place  for  the  sup- 
porting bolt.  The  center  line  may  be  drawn  if  desired  as  an 
involute  on  the  small  base  circle  shown  dotted. 

Although  the  discharge  in  the  last  two  cases  is  at  right 
angles  to  the  direction  desired  in  the  volute  casing,  the  velocity 


CENTRIFUGAL  PUMPS 


631 


has  been  reduced  to  such  a  small  value  that  even  if  this  whole 
velocity  head  were  lost  the  amount  would  be  sma1!.  More- 
over, the  head  lift  for  each  impeller  of  the  pumps  using  diffusers 
is  so  great  that  the  percentage  loss  due  to  this  radial  outflow 
is  very  small.  For  this  reason  the  volute  casings  in  these  pumps 
are  not  made  as  volutes  but  are  concentric  circular  paths 


FIG.  477. — Diffuser. 

uniting  at  the  top  of  the  pump  as  shown  in  Figs.  437,  438,  and 
441. 

In    the    problem    considered    assume   Z)d  = 


2.23x144 


;— =5.3  ft.  per  sec. 


1.40(71X22-6X4) 

If  bd  is  increased  to  ij&2  or  2.00"  the  velocity  is  changed  to 
vd=3-5  ft.  per  sec. 


632  PUMPING  MACHINERY 

If  this  velocity  is  entirely  lost  in  impact  the  loss  is 

Ld=V-fg=o.3i  ft., 

or  less  than  J  of  i%  of  the  head  per  stage. 

10.  Volute  Casing.     The  volute  casing  of  ordinary  volute 
pumps  is  designed  so  that   the  velocity  of  the  water  is  the 


FIG.  478.  —  Volute  Casing. 


same   at   all   portions.     The   quantity   coming  off  from   unit 
length  of  whirlpool  chamber  is 


The  quantity  passing  at  any  point  at  the  angle  6  from  A, 
Fig.  478,  is 


CENTRIFUGAL   PUMPS 


633 


If  the  velocity  of  this  water  is  assumed  as  vv,  the  area  at 
the  point  is 

A  =  — — = — —  (if  of  circular  section). 


If  A  is  rectangular  and  of    constant  width  w  in  direction 
of  the  axis,  the  dimension  in  the  direction  of  the  radius  is 


Vv2nw 


KO. 


/->—.«!        ftftt        I 


2nvvWA 


Neumann  points  out  that  the  first  assumption  of  a  circular 
section  gives  the  limiting  curve  of  Fig.  478  the  form  of  a  para- 
bolic spiral,  while  in  the  second  case  the  limiting  curve  is  an 
involute. 

ii.  Shaft  Design.  The  weight  of  the  impellers  should  be 
computed  from  the  drawing  made,  as  in  Fig.  464,  and  from 
previous  experience  the  length  of  the  shaft  to  care  for  these 
impellers  is  known.  A  diagram,  such  as  shown  in  Fig/479,  is 
made,  giving  weights  and  positions  between  supports. 


Bearing 


1 

135  Ibs. 
D 

135  Ibs. 

135  Ibs. 
B 

Bearing    Coup 

i 

o"      •> 

tf 

>!  i 

! 

FIG.  479. — Load  Diagram. 

Considering  the  beam  as  a  simple  beam,  the  bending  moment 
at  the  point  A  is  o  and  the  twisting  moment  T  is 


100X33,000X12  . 

T= —=6300  in. -Ibs., 

1000  X27T 


from  B  to  C, 


The  reactions  due  to  the  loads   of   135   pounds    at   each 
impeller  will  be  229 £  pounds  at  the  left  and  175^  pounds  at 


G34  PUMPIXG   MACHINERY 

the  right.  The  shear  diagram  will  pass  through  zero  at  the 
second  load,  hence  the  moment  will  be  a  maximum  at  this 
point. 

M  =  229^X13  -135X9  -=1768  in.-lbs. 

Under  the  last  impeller  the  moment  is 

M  =  175!- X8  =  1404  in.-^bs. 
The  combined  moment  is 


At  B  this  is 


Tc  =  \  14042  +  ( 6300  )2  =  6400 ; 
at  C, 


+  (  4200)2=  4550. 
For  the  shaft  diameter, 

C7T^3 

6400=^5^ 


As  will  be  seen,  the  critical  speed  for  such  a  combination  of 
discs  on  a  shaft  demands  a  larger  diameter,  and  for  that  reason 
a  diameter  of  2  has  been  assumed  before  the  investigation  for 
critical  speed. 

The  thrust  bearing  will  be  investigated  at  this  point.  The 
thrust  is  caused  by  the  impact  and  pressure  on  the  inlet  area 
of  the  pump  and  the  unbalanced  pressure  on  the  shrouding  or 
sides  of  the  impellers.  Using  the  dimensions  from  Fig.  464, 

Diameter  at  hub  =3J  ins. 

"  •         inner  clearance  ring  =8^  ins. 

"  "  back  =  7  ins. 

44  discharge  =14.6  ins. 

"          outer  edge  of  entrance      --=7^  ins. 


CENTRIFUGAL  PUMPS  635 


5^2=63.8  ft.  per  sec. 

^1=36.4  ft.  per  sec. 

c2  =  6.2  ft.  per  sec. 

Ci  =37.8  ft.  per  sec. 
24  =  604  ft-  Per  sec- 
0  =  9.80  ft.  per  sec. 


2]=62.5  ft-  head. 

If  the  water  leaks  on  each  side  to  the  center,  as  seen  in 
Fig.  419,  this  pressure  difference  acts  on  each  side  and  con- 
sequently there  is  no  unbalanced  pressure  from  this  source. 
If  there  was  leakage  only  on  one  side,  say,  the  back  of  the 
impeller,  then  it  would  be  assumed  that  the  pressure  in  space 
would  be  a  portion  of  the  62.  5.  feet,  as  the  water  leaks  out  at 
center  into  the  suction.  Suppose  this  be  assumed  to  be  two- 
thirds  of  62.5,  or  40  feet  approximately. 

The  pressure  to  the  right  is  then 

40x62.5!"  .I4.62        72  ] 
p=x       --  U--      -TT--    =2180  Ibs. 
144     L      4  4  J 

The  force  from  the  impact  of  the  water  to  the  left  is 


. 

Wv    wAc2         J4V/8 
—  =  -  =  -    -  =  50  Ibs. 
g         g  144X32.2 

If  there  are  openings  at  the  center  of  the  impeller  to  relieve 
the  pressure  and  if  leakage  occurs  on  each  side,  the  impact  is 
all  that  has  to  be  cared  for. 

The  total  pressure  to  be  carried  on  the  thrust  bearing  in  this 
pump  is  .  . 

XP=3X5o  =  150  Ibs. 
This  would  require  the  area, 
P     150 


636  PUMPING  MACHINERY 

If  there  are  four  collars  on  the  2-inch  shaft  the  approximate 
height  of  the  collars  to  give  the  proper  area  will  be 

;.        A  _J i;. 

«*D    4X*X3 

The  collars  will  be  made  ^  inch  high  for  easier  machine 

work. 

CRITICAL  SPEED 

If  a  shaft  which  is  deflected  slightly  is  caused  to  revolve, 
it  is  subject  to  centrifugal  force  due  to  the  weights  turning  at 
the  angular  speed  a>.  When  the  speed  is  low  the  shaft  has  a 
chance  to  bend  to  the  elastic  curve  due  to  the  weights  of  the 
body  and  the  various  parts  will  rotate  about  their  figure  centers, 
or  geometrical  centers,  which  are  about  the  same.  When,  how- 
ever, the  speed  is  increased  this  bending  cannot  occur  so  rapidly 
and  the  shaft,  pulleys  and  weights  may  be  assumed  to  rotate 
around  the  axis  of  the  bearings.  Following  Reynolds'  method 
as  given  by  Stanley  Dunkerley  in  his  excellent  paper  published 
in  the  Philosophical  Transactions  for  1894,  Part  A,  assume 
the  shaft  and  disc  as  shown  in  Fig.  480.  The  load  per  element 


FIG.  480.— Whirling  Shaft  and  Pulley. 

of  length  of  the  shaft  when  turning  with  the  angular  velocity 
w  at  which  the  weight  effect,  but  not  the  mass  effect,  can  be 

neglected  is 

Awdx 
Load  = (o2y  =  Lax, 

o 

where  A  =area  of  shaft  in  sq.ft. 

w=  weight  of  i  cu.ft. 
g= acceleration  of  gravity. 
y=  deflect  ion. 

Now  Shear  =^Ldx  =  V, 

where  L  =  load  per  foot  and  V  =  shear. 


CENTRIFUGAL  PUMPS  637 

Moment  =  (  Vdx  =M,  but  M =EI-~  where  E  equals  modulus 
of  elasticity  in  pounds  per  sq.ft.  and  /  is  the  moment  of  inertia 
of  the  cross-section  in  feet4.  Hence  V  = 


,  . 

dx3  dx* 

The  equation  of  the  elastic  curve  then,  when  the  speed  is  such 
that  the  deflection  is  caused  by  the  centrifugal  force  and  not 
by  the  weight,  is 

Aw  ,  d4y 

—co2y=EI^L. 
g  dx4 

This  occurs  when  the  effect  of  weight  is  eliminated  or  the 
shaft  is  moving  at  such  a  speed  that  whirling  occurs. 
The  equation  may  be  written  as 


where  m= 


. 
gEI 

This  integrates  into 

y  =A'emx  +B'e~mx  +  C'emix  +D'e~mix, 
or  y  =A  cosh  mx  +B  sinh  mx  +  C  cos  mx  +D  sin  mx. 

A,  B,  C,  D  are  the  constants  of  integration. 

To  eliminate  these  constants  there  are  several  known  con- 
ditions.    At  bearings,  x  =  o  or  /,  and  y  =  o;    for  fixed  bearings 

•j  ""P,  at  a  change  of  loading  x,  y,  and  -=-  for  the  equations 

on  one  side  of  the  load  are  equal  to  the  respective  equations 
which  hold  on  the  other  side  of  this  point.  The  equation 

above  has  been  derived  from  the  condition  that  L=  —  o>2y 

6 

for  the  curve  and  this  only  holds  between  concentrated  loads. 
At  each  concentrated  load  or  bearing  for  a  shaft  with  several 
bearings  there  is  a  new  equation.  At  the  concentrated  loads 


638 


PUMPING  MACHINERY 


When  there  is  a  disc  on  the  wheel  which  is  deflected  by 
the  bending  of  the  shaft  into  the  position  shown  in  Fig.  481 
the  centrifugal  action  of  the  disc  tends  to  right  the  disc  and 
straighten  the  shaft. 

For  an  element  dm  of  the  disc  at  distance  r  from  the  center, 
of  which  the  components  are  v,  t  and  6,  the  component  of  the 

dm 
centrifugal  force  tending  to  right  the  disc  is  — vco2  and  its  arm 

o 

is  t.     The  moment  of  this  is  tvco2 —  or  -^-v2a? — , 

g       dx       8 

t     dy 
since  —=-T- 

.     v     dx 


FIG.  481.—  Rotating  Disc. 

The  total  righting  moment  is 


j 

g  d 


g  dx 

y  f 

I  v2dm. 
xj 


Now 
hence 

//  is  very  small, 


—Ip  =  l'd  (about  diameter). 


CENTRIFUGAL   PUMPS  639 

hence  M,  the  moment  due  to  centrifugal  force  tending  to  right 
the  disc,  is 


M  =         !'. 
g  ax 

In  passing  a  disc  the  difference  between  the  moments  on  the 
two  sides  is 


This  is  another  condition  in  passing  a  load  which  elim- 
inates another  constant.  These  conditions  will  furnish  suffi- 
cient equations  to  eliminate  the  constants. 

Dunkerley  in  his  extensive  article  takes  up  the  different 
cases  arising  in  practice  and  determines  the  critical  speed. 
The  student  is  referred  to  the  article  for  the  methods  of  solving 
the  equations,  but  the  simple  case  of  a  plain  shaft  is  given 
here  to  show  how  these  are  worked  out. 

For  a  simple  shaft,  the  equation  for  y  is 

y  =A  cosh  mx  +B  sinh  mx  +  C  cos  mx  +D  sin  mx. 

Now  when 

*  =  oor  /,  y=o, 
and  when 

d2y 
x  =  o  or 7,  M=o,  i.e.,  ~j~^>=°- 

-j-  =mA  sinh  mx+mB  cosh  mx  -mC  sin  mx  +mD  cos  mx. 
dx 

d2v 

— -  =m2A  cosh  mx-\-m2B  sinh  mx  —m2C  cos  mx  —m2D  sin  mx. 

dx2 

0=A+C. 

0=A  cosh  ml+B  sinh  ml  +  C  cos  ml+D  sin  ml. 
0=A-C. 

0=A  cosh  ml+B  sinh  ml  -  Ccos  ml—D  sin  ml. 
:.  A  =o,  C=o. 

B  sinh  ml+D  sin  ml  =  o. 

B  sinh  ml—D  sin  ml=o. 


Subtracting  Dsmml=c. 

D=o  or  ml 


640  PUMPING  MACHINERY 

Adding 

B  sinh   ml=o. 
B=o  or  ml  =  o. 

The  condition  which  is  possible  is 

ml  =  n, 

£7/4  =  ?r4' 

W  ==  ~jZt'\  I     j       ==          » 

I2  \  Aw     30 


In  the  case  of  a  disc  on  a  shaft  as  shown  in  Fig.  481  the 
following  results  if  the  mass  of  the  shaft  is  not  considered, 
although  the  resistance  to  deflection  is  taken  into  account : 


=, 

dx4 

EIy~x3+-- 

O  2 

x=o,  y=o. 

d2y 


__ 
dx=dx~/' 


tf~dx~      El  g  dx 
/=o. 

v' 

=M     =o. 


CENTRIFUGAL  PUMPS  641 

These  give  the  following  equations  in  order 


Ac3  A'c*     B'c2 

+  -  +c'c+D'. 
62 


is  the  square  of  the  radius  of  gyration      -. 


0=A'l+B'. 

Eliminating  the  constants  A,  B,  C,  D,  A',  B',  C',  D',  from 
these  equations  Dunkerley  obtains  the  following  equation, 
calling  l—c=c'\ 


Ji[_3L     Tl          3/T  _ 
acc'U^c'2  acc'L 


3/ 


3/ 

but  / 


K*=  ^ti-ay 

•tx  _-0-/o         / -/ 


Now  ^T2  must  be  a  positive  quantity,  since  it  equals 

Hence 

ac2c'2 
K-—T- 


642  PUMPING  MACHINERY 

w 

Now   a=—pj.aj2   and   these   inequalities   give   the  limiting 

speeds  to  for  the  centrifugal  force  to  act  to  produce  whirling. 
These  are  only  the  limiting  values  beyond  which  there  is 
certainly  no  whirling  or  beyond  which  it  is  known  to  be  with- 
out the  range  of  the  critical  speed.  To  get  the  critical  speed 
the  equation  for  K2  is  solved  for  a  and  then  a  is  expressed  in 
terms  of  w2,  giving 


4«2b 
i-6)3J' 


In  this  fl=^=ratio  of  the  distance  from  pulley  to  nearer 

bearing  to  the  rectangular  radius  of  gyra- 
tion. 

6=y=ratio  of  distance   from  pulley  to  bearing  to 

the  span  (less  than  J). 
Calling  |  of  the  bracket  above  62,  Dunkerley  gives 


We3' 


In  this  equation  /  is  the  moment  of  inertia  of  the  shaft 
section,  W  is  the  weight  of  the  disc  or  pulley,  c  is  the  distance 
from  the  nearer  bearing  to  the  disc,  6  is  a  factor  which  depends 

on  the  ratio  y  and  on  the  ratio  -^  where  K  is  the  rectangular 

l>  x\. 

radius  of  gyration. 

Dunkerley  then  computes  a  table  for  6  for  various  values 
of  a  and  b  which  has  been  reproduced  in  the  form  of  a  series 
of  curves  shown  in  Fig.  482. 

Dunkerley  considers  the  case  of  a  shaft  with  three  supports 
and  also  the  case  of  a  span  with  an  overhanging  end  of  length 
c.  If  the  ratio  of  the  short  span  to  the  longer  span  is 


CENTRIFUGAL   PUMPS 


643 


oii\-j=a]  and  if  the  same  symbol  a   be  used  for  the  ratio  y, 

for  the  second  case,  he  gives  the  specific  speed  for  each  in  the 
form 


l=K. 


In  these  two  cases  the  values  of  K  depend  on  OL  and 
Dunkerley  has  computed  the  values  of  K  for  different  values 
of  a.  These  have  been  plotted  in  Fig.  482  so  that  CD  may  be 
found  and  from  it  N. 


0.25 


0.50 


0.75 


1.00 


1.50 


1.75 


200 


FIG.  482.  —  Dunkerley's  Values  for  Shafts. 

For  the  case  of  a  disc  on  the  projecting  end  of  length  c, 
the  formula 


holds  for  the  disc  independent  of  the  shaft. 

The  value  of  0,  as  in  the  case  of  the  disc  between  supports 

c  c 

depends  on  the  value  of  a=^  and  b=-j.     Dunkerley's  values 


644  PUMPING  MACHINERY 

of  6  for  four  values  of  j  =b  have  been  plotted  for  different 

values  of  a  =-^.     The  value  of  0  may  be  taken  from  the  curves 
A 

from  which  aj  and  N  may  be  determined. 

In  the  case  of  the  double-span  shaft  it  is  seen  that  the 
value  of  K  does  not  vary  much  from  TT,  showing  that  the  main 
effect  of  using  several  bearings  is  to  shorten  the  span  only. 
This  shortening  has  considerable  effect,  although  the  continuous 
beam  action  is  not  important.  If  there  must  be  an  overhanging 
end  of  a  shaft  of  fixed  total  length  it  may  be  shown  that  when 

-=£  or  the  bearing  is  £  the  total  length  from  the  overhanging 
If 

end  the  critical  speed  will  be  the  greatest. 

With  the  curves  of  Fig.  482  the  critical  speed  can  be  figured 
for  each  disc,  wheel  or  pulley  on  a  shaft  independently  of 
each  other  and  of  the  shaft.  When  these  are  computed  the 
resultant  speed  is  such  that 

JL    _* L.  ,  JL  , 

Nr2    Ni*^Nf-N#^ 

where  Nr  is  the  resultant  speed  and  NI,  N2,  N3,  etc.,  are  the 
critical  speeds  due  to  any  one  part  independent  of  the  other 
parts.  Hence 

N,N2N3N4 .  . . 


...Nn2+N12N23...Nn2+  ...' 

or  for  the  case  in  hand  of  three  impellers  and  a  shaft,  suppose 
NI  is  the  critical  speed  for  the  shaft,  and  N2,  N3,  and  N± 
those  for  the  impellers.  The  critical  speed  of  the  combina- 
tion will  be 


N  = 


If  these  happened  to    be  identical,   the  following  would 
result: 


CENTRIFUGAL  PUMPS 


645 


In  the  design  of  the  shaft  the  stuffing  boxes  are  con- 
sidered to  be  the  same  as  bearings  on  account  of  the  closeness 
of  fit.  This  gives  for  investigation  of  critical  speed  a  shaft 
of  three  spans  with  an  overhung  disc.  The  shaft  will  be  con- 
sidered alone  as  a  shaft  with  a  1 2-inch  span  and  a  3o-inch 
span,  as  no  constants  are  given  for  shaft  of  three  spans  such 
as  is  shown.  This  gives  the  lowest  speed  of  any  combina- 
tion for  the  shaft.  All  dimensions  are  in  feet. 


7T(2) 


E  =  29,000,000  X  144. 
12 


4-^ft*. 

I24 


30 


0.4. 


_30#2   fJEI 
1  *~    Til2  \Aw 


0.785 
32  X  29,000,000  X 144  X7^4 

22X46o 


4X144 


For  the  impeller  discs  the  weights  are   taken  as  90  pounds 

D 

each  and  the  radius  of   gyration  =—.  The  separate  discs  give 

the  following: 


Disci,  61=— =0.13,     0!  =--  =  !. i,       #1=1.05. 


Disc  2,  62  =  ^=0.43, 


3.65 


=  3.56,      02=2.1. 


Disc  3,  63  = — =0.26,     03=— £-=2.2,       03=o.o. 
3°  3-65 


30X1.05     /32  X  29,000,000  X  0.785  _ 
Wcs       -^-        ~  /.A  a  =12,400. 


-J  xx^s 


646  PUMPING   MACHINERY 


3 

=458a 


3760.          ,      .  .  . 

For  the  disc  on  the  end  #  =  -  =  1.5,  b=  —  =J,  #  =  --=4, 
2.o,  c  =6,  W  =  i5lbs. 


12 


2.0     //4\3/9°\ 
,400  ---  A/   z~  =  31>3oo« 

i.o5\\6/  \i5/ 


The   resultant   of    these   various   speeds    is   given   by   the 
formula 


The  speed  N3  is  the  factor  which  has  the  greatest  control 
of  the  resultant.  These  terms  depend  on  \  /  for  their  values, 
so  that  if  the  diameter  of  the  shaft  was  reduced  to  ij,  the 
approximate  value  of  N  would  be 

~T 

I680, 

a  value  so  close  to  the  actual  speed  that  there  might  be  con- 
siderable vibration  and  shock.  This  makes  clear  the  reason 
for  the  increase  in  the  diameter. 

CENTRIFUGAL  PUMPS  FOR  SPECIAL  PURPOSES 

The  centrifugal  pump  has  been  used  for  many  years  for 
the  clearing  of  dry  docks.  Fig.  483  illustrates  the  equip- 
ment for  one  of  the  docks  at  the  League  Island  Navy  Yard. 
The  units  are  45-inch  volute  pumps  built  by  Worthington; 
each  is  driven  by  a  450-!!. P.  motor  and  will  handle  an  average 
quantity  of  50,000  gallons  per  minute  against  heads  varying 


CENTRIFUGAL   PUMPS 


647 


from  i  foot  to  33  feet.  As  was  pointed  out  with  the  test 
curves,  the  quantity  decreases  as  the  head  increases,  the  speed 
remaining  constant.  The  power  usually  reaches  a  maximum 
at  an  intermediate  head,  so  that  there  is  no  danger  of  over- 
loading the  motor  in  any  case.  It  is  this  feature  which  makes 
the  centrifugal  pump  of  value  for  this  service. 

The  quantity  of  50,000  gallons  per  minute,  which,  is  the 
usual  way  of  rating  centrifugal  pumps,  would  mean  72,000,000 
gallons  per  twenty-four  hours,  the  method  used  in  rating 
water- works  pumps. 


FIG.  483.— Dry  Dock  Units.     (Worthing ton.) 

These  pumps  for  many  years  have  been  used  extensively 
for  this  work  and  for  draining  the  low  lands  of  Holland  and  the 
marshes  of  Italy.  Fig.  484  shows  one  of  the  stations  built  by 
Gwynne  in  1876.  There  are  eight  pumps  placed  in  pairs 
driven  by  compound  engines.  The  boiler  houses  are  located 
at  the  ends  of  the  pump  room.  This  arrangement  shows  a 
good  plan  and  one  well  thought  out.  The  pumps  were  of  double- 
flow  volute  type  handling  57,000  gallons  per  minute  under  a 
head  of  7}  feet.  The  impellers  were  60  inches  in  diameter  and 
two  of  them  were  driven  by  a  27f-inch  and  46! -inch  by  27-inch 
engine.  The  plant  as  shown  in  Engineering,  was  built  for  the 
Ferrara  Marshes  in  northern  Italy. 

The  pumps  of  that  day  gave  efficiencies  of  50  to  70. per  cent 


648 


PUMPING  MACHINERY 


CENTRIFUGAL  PUMPS 


649 


and  were  very  reliable.  The  higher  efficiencies  obtained  to-day 
are  due  to  the  better  method  of  design. 

One  of  the  largest  centrifugal  pumps  built  in  1884  by  Simp- 
son &  Co.  for  the  London  docks  of  the  East  and  West  India 
Companies,  handled  46,400  gallons  per  minute. 

The  pumps  shown  in  Fig.  485  are  remarkable  in  that  they 
handle  35,000  gallons  per  minute  under  the  high  head  of  160 
feet,  requiring  a  2ooo-H.P.  motor  to  drive  each  of  them.  These 
pumps  were  of  the  turbine  type  on  account  of  the  high  head 


FIG.  485. — 36-Inch  Turbine  Pump  of  Worthington. 

and  were  used  at  the  St.  Louis  exposition  to  supply  water 
to  the  Grand  Cascade.  The  volute  casing  of  concentric  form 
with  an  outlet  at  right  angles  is  common  with  this  type  of 
pump.  The  figures  show  how  part  of  the  diffuser  is  cast  with 
the  volute  casing  on  one  side  while  the  head  containing  the 
suction  flange  and  elbow  forms  the  other  part  of  the  diffuser. 
Fig.  486  illustrates  a  casing  of  an  R.  D.  Wood  pump 
of  50,000  gallons  per  minute  for  condenser  work.  This  is  a 
45-inch  pump.  Many  pumps  of  this  type  are  used  with  verti- 
cal axes.  The  figure  illustrates  the  method  used  in  forming  the 
casings  of  pumps  of  large  diameter.  Much  ingenuity  is  dis- 
played at  times  in  the  methods  of  separating  the  parts. 


650 


PUMPING  MACHINERY 


The  pump  shown  in  Fig.  487  is  one  recently  built  by  the 
Alberger  Company  for  the  Standard  Oil  Company.  It  is  a 
turbine  pump  and  consists  of  three  units  connected  in  parallel, 
as  was  the  case  in  the  tri-rotor  volute  pump.  This  pump  is 
to  handle  20,000,000  gallons  per  twenty-four  hours.  The  form 
of  the  casing  with  the  discharge  outlets  at  right  angles  to 
the  main  direction  of  flow  is  seen  here  as  in  the  other  turbine 
pumps.  . 


FIG.  486.— R.  D.  Wood  45-Inch  Centrifugal  Pump. 

Another  important  use  for  the  centrifugal  pump  is  that  of 
dredging  channels.  In  this  case,  silt,  sand  and  even  rocks  are 
raised  with  the  water  by  the  centrifugal  pump  and  delivered 
into  scows  or  settling  basins  on  land  through  long  flexible 
pipes  carried  on  pontoons.  The  solid  material  settles  out 
from  the  water  and  the  water  is  returned  to  the  stream.  This 
method  forms  a  very  cheap  and  effective  manner  of  dredging 
where  possible.  The  pump  has  to  be  so  constructed  that 
the  blades  of  the  impeller  may  be  easily  replaced  when  broken 
or  worn  out.  It  is  quite  evident  that  the  solid  matter  will 


CENTRIFUGAL   PUMPS 


651 


cause  rapid  wear  of  these  parts.  It  is  also  important  to  have 
all  passages  direct  and  of  ample  size  to  pass  large  pieces  of 
solid  matter  which  may  be  carried  through  the  pump. 

Mr.  F.  B.  Malt  by,  in  the  Transactions  of  the  American 
Society  of  Civil  Engineers,  Vol.  54,  gives  an  excellent  descrip- 
tion of  dredges  used  on  the  Mississippi  River  and  the  machinery 
on  them.  Mr.  Geo.  Fowler  gives  a  drawing  of  the  type  used 
in  New  York  harbor  in  Vol.  31,  p.  468,  of  the  Transactions  of 


FIG.  487. — 20,000,000  Gallon  Alberger  Multi-impeller  Turbine  Pump. 

the  American  Society  of  Mechanical  Engineers.  From  his 
description  Fig.  488  has  been  prepared.  The  figure  shows 
the  method  of  attaching  the  vanes,  tips  and  the  large  passages 
used  with  these  pumps.  Mr.  Fowler  reports  that  in  dredging 
the  New  York  Ship  Channel  a  piece  of  shaft  weighing  70 
pounds  was  lifted  and  passed  by  a  centrifugal  pump,  and  at 
Yonkers,  N.  Y.,  an  8-inch  pump  on  a  wrecking  boat  lifted 
and  passed  a  35 -pound  piece  of  pig  iron  u^X4f  X3i  inches. 
These  dredge  pumps  are  driven  at  a  speed  of  about  noo  feet 
per  minute  at  178  R.P.M.  Those  used  in  New  York  harbor  on 


652 


PUMPING   MACHINERY 


one  contract  were  driven  by  iga-H.P.  engines,  delivering 
10,000  gallons  per  minute.  Mr.  Fowler  shows  the  detail  of 
the  end  of  the  suction  line.  This  end  piece  sinks  in  the 
soft  silt  or  sand  and  the  solid  matter  is  lifted  with  the  water. 
Extra  openings  permit  water  entering  the  suction  pipe  when 
the  silt  covers  the  mouth  too  much  for  proper  action. 


FIG.  488. — Dredging  Pump. 

One  of  the  latest  uses  for  centrifugal  and  piston  pumps 
is  that  for  the  high-pressure  fire  service.  The  great  fire  hazards 
in  the  congested  districts  of  trade  in  our  large  cities  has  made  it 
necessary  to  build  a  separate  high-pressure  water  supply  system 
throughout  these  districts.  The  pumping  stations  may  be 
placed  on  a  river  front,  where  an  unlimited  supply  of  water 
may  be  had,  or  in  the  center  of  the  city,  in  which  case  a  special 
supply  line  of  large  size  is  brought  from  one  of  the  city  reser- 
voirs. The  use  of  salt  water  in  cities  such  as  New  York  or 


CENTRIFUGAL   PUMPS 


653 


Doo 


654 


PUMPING  MACHINERY 


Brooklyn  is  objected  to  on  account  of  the  damage  done  by 
salt  water  when  it  touches  merchandise.  For  this  reason  in 
New  York  the  supply  is  taken  from  the  fresh- water  mains. 
The  use  of  water  for  fire  service  in  such  a  city  is  a  very  small 
percentage  of  the  total  water  used,  so  that  this  is  not  a  very 
expensive  method,  and  in  the  usual  system  employing  fire 
engines  the  water  is  drawn  from  fire  hydrants  on  the  fresh 
water  supply,  so  that  the  cost  of  water  is  not  increased  when 
city  water  is  used  with  the  new  stations. 


FIG.  490. — Philadelphia  High-pressure  Station. 

From  the  pumping  stations  extra  heavy  flanged  piping  is 
carried  in  a  network  over  the  district.  Special  hydrants  are 
used  and  in  many  cases  fire  lines  are  led  into  the  various 
buildings.  Valves  must  be  used  at  frequent  intervals  to  con- 
trol lines  leading  into  buildings  or  sections  of  the  system,  so 
that  the  waste  of  water  could  be  prevented  in  case  a  wall 
should  fall  covering  the  valve  controlling  the  branch  leading 
to  a  burning  building. 

When  a  fire  occurs  the  alarm  is  sent  to  the  pumping 
station  and  the  pumps  are  put  into  commission.  One  of  the 
earliest  of  these  stations  was  erected  for  the  city  of  Philadel- 
phia. The  pumps  were  of  the  triplex  form  driven  by  gas 
engines.  These  are  shown  in  Fig.  489.  The  gas  engines  are 
furnished  with  gas  from  a  special  city  main.  The  engines 


CENTRIFUGAL   PUMPS 


655 


are  started  by  compressed  air  held  in  storage  tanks  at  one 
end  of  the  building.  There  are  several  methods  of  igniting 
the  charge,  so  that  should  one  method  fail  another  may  be 
employed.  The  Westinghouse  engines  used  in  this  station  have 
always  responded  to  the  demands  of  the  service.  The  Dean 
Triplex  Pumps  were  built  to  give  1200  gallons  per  minute  under 
300  pounds  pressure.  The  gas  engines  gave  280  H.P.  each. 

Fig.  490  shows  the  seven  I2oo-gallon  double-acting  pumps 
and  the  two  35o-gallon  pumps  originally  put  in  with  the  two  air 


FIG.  491. — High-pressure  Pumping  Station,  Brooklyn,  N.  Y. 

compressors  placed  at  the  west  end  of  the  building.  These 
air  compressors  are  used  to  charge  the  air  tanks,  which  are 
made  up  of  heavy  piping.  The  suction  pipe  enters  from  the 
river,  passing  between  the  two  rows  of  pumps.  It  is  connected 
to  the  various  pumps  by  motor-operated  valves.  The  pumps 
are  started  under  no  load  and  after  the  engine  is  operating 
properly  the  by-pass  valve  on  the  pump  is  closed,  allowing 
the  discharge  to  enter  the  pressure  main.  The  pressure  main 
is  connected  by  a  check  valve  with  the  city  mains,  so  that  there 
is  always  a  pressure  of  60  pounds  in  this  line.  Provisions 
are  made  for  connecting  fireboats  to  the  pressure  line  in  case. 


656 


PUMPING   MACHINERY 


there  should  be  a  need  for  it.     The  gas  engine  is  off  centered 
from  the  pump  to  accommodate  the  gears. 

In  1906  the  same  kind  of  system  was  introduced  in 
Brooklyn,  using  turbine  pumps  driven  by.  means  of  electric 
motors.  The  Brooklyn  station  is  shown  in  Fig.  491.  The  five- 


PLAN  OF  THE  STATIONS,  SHOWING  PUMPS  AND  PIPING 

FIG.  492. — New  York  High-pressure  Station. 

stage  Worthington  turbine  pumps  deliver  3000  gallons  per 
minute  under  a  head  of  300  pounds.  The  test  curve  from 
these  pumps  is  shown  in  Fig.  458.  The  use  of  electric  motors 
for  such  stations  must  be  safeguarded  by  the  use  of  several 
cross-connected  power  houses,  so  that  an  accident  to  one  will 
not  endanger  the  reliability  of  the  station. 


CENTRIFUGAL   PUMPS 


657 


One  of  the  latest  installations  is  that  for  New  York  city.  There 
are  two  stations,  one  on  East  River  at  Oliver  Street  and  one  on 
the  North  River  at  Gansevoort  Street.  Each  station  has  five  five- 
stage  Allis-Chalmers  centrifugal  pumps  driven  by  Allis-Chalmers 
induction  motors.  The  plan  of  the  station  is  shown  in  Fig. 

492,  while  the  interior  of  one  of  the  stations  is  shown  in  Fig. 

493.  This  photograph  shows   the   appearance  of  the  pump 


FIG.  493. — New  York  High-pressure  Station. 

with  its  outflow  casing  at  the  left  hand  end  and  the  valve  on  the 
suction  pipe  at  the  right.  The  3-phase,  25-cycle,  63oo-volt, 
74O-R.P.M.  induction  motors  are  also  seen.  The  pressure  gauges 
give  some  idea  of  the  proper  operation  of  the  pumps.  The 
extended  thrust  bearing  with  its  cooling  pipes  is  seen  at  the* 
left.  A  section  of  this  pump  was  given  in  Fig.  454. 

There  are  two  suction  pipes  leading  to  the  river  at  the 
top  of  Fig.  492.  And  in  addition  there  are  two  fresh- water 
suction  pipes  entering  from  the  city  mains  on  the  sides. 
The  two  air  chambers  on  the  river  suction  pipes  are  for  the 


658  PUMPING  MACHINERY 

purpose  of  steadying  the  flow.  These  are  kept  charged  by  the 
suction  air  pump  shown  at  the  upper  right-hand  corner.  The 
suction  pipes  form  a  loop  around  the  station.  The  discharge 
mains  are  also  arranged  in  a  loop  and  both  lines  are  equipped 
with  valves,  so  that  any  section  of  the  pipe  may  be  cut  out 
without  affecting  the  operation  of  the  station.  Venturi  meters 
are  used  on  the  discharge  to  measure  the  water  used  by  the 
system. 

The  discharge  from  each  pump  is  controlled  by  a  valve, 
so  that  as  soon  as  the  pressure  in  the  main  becomes  greater 
than  the  predetermined  amount,  due  to  the  decreased  use  of 
water,  a  special  valve  by-passes  a  portion  of  the  discharge  into 
the  suction.  In  all  of  the  central  .high-pressure  stations  the 
pressure  is  regulated  according  to  the  wishes  of  the  fire  .chief, 
as  he  is  in  telephonic  communication  with  the  station  at  all 
times.  The  pressure  is  changed  by  changing  the  number  of 
units  in  operation. 


CHAPTER  XV 

MINE    PUMPS 

THE  method  of  mine  pumping  has  been  changed  very  mate- 
rially of  late  years.  The  early  method  introduced  by  Newcomen 
is  the  one  which  has  been  usually  followed  until  recent  times. 
Fig.  494  shows  the  method  of  using  a  long  rod  extending  from 
the  engine  house  to  the  pump  barrels.  When  necessary  to 
take  off  a  pump  in  a  side  gallery  a  bell-crank  lever  was  mounted 
at  the  side  of  the  rod,  one  end  of  which  was  attached  to  the 
rod;  the  other,  to  the  branch  rod  or  pump. 

When  necessary  to  balance  these  long  rods,  beams  were 
mounted  above  ground  or  in  the  rod  shaft  and  on  these  counter- 
balance weights  were  attached. 

By  arranging  the  bell  crank  lever  so  that  its  pump  operates 
on  the  stroke  of  the  main  pump  on  which  no  pumping  is  done, 
balancing  may  be  accomplished. 

These  reduced  the  load  to  be  moved  by  the  piston,  but  the 
inertia  of  the  system  was  greatly  increased.  Another  method 
used  at  times  was  to  place  two  pumps  in  the  same  shaft, 
balancing  one  set  of  rods  by  the  other.  Such  a  pump, 
built  in  Aix-la-Chapelle  (Fig.  495),  illustrates  this  method 
very  clearly  and  gives  some  idea  of  the  complicated  system 
used.  The  figures  illustrate  engines  with  fly  wheels  although 
many  are  used  without  fly  wheels,  direct-acting  connections 
being  employed.  Some  of  these  pumps  used  in  America  are 
remarkable  for  their  size  and  weight.  The  Mexican  Union 
Pump  of  1880,  built  with  Leavitt  jacketed  cylinders,  had  64X96 
inches  for  H.P.  cylinders  and  100X102  for  L.P.  cylinder.  A 
36-foot  fly  wheel  was  used.  The  pumps  were  arranged  at  various 
points  of  the  rods,  there  being  fourteen  different  plungers. 
The  rods  were  2618  feet  long,  and  there  was  so  much  con- 

659 


660 


PUMPING  MACHINERY 


FIG.  494. — Mine  Pump. 


MINE   PUMPS 


661 


traction  in  the  rods  during  action  that  the  lo-foot  stroke  at  the 
engine  was  reduced  to  97  inches  at  the  pumps.     This  system  of 


using  a  number  of  pumps  in  the  total 
lift,  in  this  case  1180  feet,  is  often 
used  in  this  type  of  mine  pump. 
The  use  of  these  reduces  the  amount 
of  head  on  the  various  plunger  barrels 
and  makes  it  possible  to  use  lighter 
parts. 

In  the  Mexican  Union  Pump  the 
total  weight  of  the  moving  parts  was 
1,620,500  pounds.  The  use  of  a  fly 
wheel  in  this  •  pump  is  not  com- 
mon; many  pumps  do  not  have 
them. 

The  Yellow  Jacket  Shaft  Pump 
of  1880  was  one  of  the  best  pumps 
of  this  time.  It  was  a  compound 
horizontal  fly-wheel  engine  pump. 
The  steam  end  had  31-  and  62-inch 
cylinders  with  a  lo-foot  stroke.  The 
rods  were  3055  feet  long  and  the 
moving  weight  was  1,510,400  pounds. 
This  pump  made  only  5^  R.P.M. 
when  giving  its  maximum  dis- 
charge of  750  gallons  per  minute. 
A  number  of  other  pumps  in  this 

FIG.  495 ••—German  Mine  Pump,  district   were   built   with  compound 

cylinders  32X129  and  65X96  inches 

with  no  fly  wheels  and  a  total  weight  of  moving  parts   of 

1,437,900  pounds. 

At  the  Consolidated  California  Shaft,  a  Davey  Differential 


662  PUMPING   MACHINERY 

Pump  with  steam  cylinders  24X90  and  40X96,  and  pumps  of 
7  feet  stroke  lifted  500  gallons  per  minute.  The  moving  parts 
with  2150  feet  of  pump  rod  weighed  860,000  pounds. 

A  number  of  such  engines  to  handle  5400  gallons  per  minute 
against  a  head  of  1152  feet  cost  $1,300,000  without  foundations 
or  installation.  The  cost  of  operation  when  5040  gallons  were 
lifted  1074  feet  was  $58,210  per  month,  or  $677,440  per  year. 

In  contrast  to  this  cumbersome  method,  the  method  of 
installing  independent  pumps,  Fig.  496,  is  to  be  mentioned. 
In  this  case,  the  pump  is  installed  at  a  point  where  needed, 
and  by  the  use  of  proper  air  chambers,  the  effect  of  the  inertia 
in  the  water  can  be  eliminated. 

The  pump  may  be  made  sufficiently  strong  and  by  using 
the  express  type  of  pump,  or  the  centrifugal  form,  a  sufficient 
capacity  may  be  installed  in  a  small  space.  The  operation  of 
such  a  pump  must  be  by  steam,  compressed  air,  or  electricity. 
The  first  method  is  out  of  the  question  for  great  depths,  as 
the  condensation  of  steam  means  much  trouble;  moreover, 
the  radiation  losses  from  the  steam  pipe  would  be  excessive 
even  if  the  use  of  high  superheat  brought  the  steam  to  the 
pump  in  a  dry  condition.  The  use  of  compressed  air  may 
be  possible  if  a  high  grade  compressor  is  used  and  the  exhaust 
may  be  used  for  ventilation.  The  loss,  however,  in  this  case 
is  one  which  must  be  considered  carefully.  The  last  method 
of  the  use  of  the  electric  motor  is  one  which  has  many  points 
of  advantage. 

Power  may  be  brought  to  the  motor  in  a  very  simple  manner 
and  the  location  of  the  apparatus  is  not  difficult.  The  trans- 
mission loss  with  moderately  high  voltages  and  proper  frequency 
is  not  great. 

Fig.  496  shows  a  shaft  in  which  water  is  raised  from  one 
pump  at  a  great  depth  and  discharged  into  another  at  a  con- 
siderable height  above  it.  The  water  is  then  lifted  by  a  second 
pump  through  the  remaining  height.  The  express  pump 
shown  in  the  figure  with  its  motor  or  a  centrifugal  pump  uses 
a  very  small  amount  of  space  when  compared  with  the  older 
forms  of  slow-running  pumps  even  if  driven  by  steam,  air  or 


MINE  PUMPS 


663 


FIG.  496. — Express  Pumps  in  Series. 


664 


PUMPING   MACHINERY 


pump  rod.     In  this  way  considerable  expense  is  saved  in  pre- 
paring a  pump  room. 

The  cost  of  the  apparatus  in  this  case  has  been  very 
materially  reduced.  Two  i6oo-gallon  independent  high-speed 
pumps  to  work  against  1500  feet  head,  and  each  driven  by 
an  800  H.P.  slow-speed  induction  motor  cost  $80,000  and 
they  weighed  with  the  motors  600,000  pounds.  The  same 
capacity  of  rod-driven  pumps  would  cost  $960,000  and  the 


FIG.  497. — Knovvles  Express  Pump. 

moving  parts  alone  would  weigh  over  5,000,000  pounds.  The 
cost  would  be  twelve  times  and  the  weight  eight  times  as 
much  as  that  of  the  high-speed  electrically-driven  pump. 

The  express  pump  of  Riedler  was  explained  on  p.  163  and 
a  section  of  it  was  shown  in  Fig.  154.  After  Riedler  had 
shown  that  high  speeds  were  possible  with  proper  valve  and 
proper  air  chambers,  others  took  up  the  construction  of  these 
pumps  and  found  that  the  mechanically-closed  valves  were 
not  necessary  and  that  self-closing  valves  could  be  used.  In 
some  cases  the  valves  are  of  small  diameter. 


MINE  PUMPS 


665 


Fig.  497  is  a  Knowles  express  pump  intended  for  mine 
service.  The  pump  is  a  duplex  double-acting  pump,  6iJxi5 
inches  which  has  a  capacity  of  1600  gallons  per  minute,  against 
a  head  of  1550  feet.  The  motor  is  of  300  R.P.M.  This  speed 
is  very  high  and  the  inertia  of  the  water  column  is  so  great 
that  there  would  be  a  continuous  breaking  of  the  column  with 
its  accompanying  knocking  if  the  air  chambers  and  valves 


FIG.  498. — Knowles  Express  Pump. 

were  not  properly  designed.  Poppet  valves  are  the  types 
used  with  this  pump!  They  have  the  advantage  of  being 
almost  balanced  and  of  giving  a  large  area  of  discharge. 

The  figure  illustrates  the  heavy  construction  necessary  on 
the  outlet  side,  the  air  piping  for  charging  the  chambers, 
the  outside  rods  to  connect  the  front  and  back  cross-heads, 
the  pressure  oiling  system  and  the  starting  by-pass  and 
priming  valves.  The  general  lines  shown  in  the  figure  are 
excellent  and  the  design  is  massive  and  so  arranged  as  to 


666 


PUMPING  MACHINERY 


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FIG.  499. — Test  Curves  of  Express  Pump. 


MINE  PUMPS  667 

care  for  the  strains  brought  on  during  the  operation  of  the 
pump. 

The  pump  shown  in  Fig.  498  is  a  smaller  unit,  driven  by 
a  loo-H.P.  direct-current  motor.  This  pump  is  also  built 
by  the  Knowles  Steam  Pump  Works.  It  was  intended  to 
operate  at  300  R.P.M.,  pumping  250  gallons  per  minute  against 
a  head  of  1000  feet.  The  plungers  were  3^  inches  in  diameter 
and  of  5j-inch  stroke.  The  machine  is  self-contained  and 
the  frames  and  bed-plate  are  constructed  to  make  a  rigid 
structure.  The  plungers,  which  are  outside  packed,  are  con- 
nected by  outside  rods. 

A  test  of  this  pump  gave  the  results  which  are  shown  by 
curves  in  Fig.  499.  The  result  of  94  per  cent  efficency  for  the 
pump  and  84  per  cent  for  the  combined  unit,  lifting  water 
1 200  feet  is  remarkable.  The  electric  line  loss  in  operating 
such  a  pump  may  be  rendered  small  by  using  a  high  voltage, 
when  the  installation  becomes  very  efficient. 

The  types  of  direct-acting  steam  pumps  used  for  mine 
work  have  been  discussed  in  Chapter  III  and  the  forms  illus- 
trated by  several  figures.  Fig.  500  illustrates  another  one 
of  these  pumps  built  by  the  Jeanesville  Iron  Works  Co. 

It  is  intended  to  lift  1200  gallons  per  minute  against  a  head 
of  700  feet.  The  size  of  the  valve  pots,  the  control  valves 
and  gauges,  the  arms  of  the  rotating  steam  valves,  the  method 
of  valve  operation,  and  the  forms  of  the  cylinders  may  all 
be  clearly  seen. 

The  valve  pots  are  particularly  large  in  this  pump  because 
of  the  large  size  of  valves  used  by  this  company.  Fig.  501 
shows  two  forms  of  valves  used  by  them  for  heads  up  to  750 
feet.  The  annular  valve  is  used  when  the  head  is  not  greater 
than  1000  feet.  This  figure  illustrates  the  method  of  lining 
the  valve  pot  with  wood  when  the  water  contains  acids  which 
will  attack  the  metal. 

The  design  of  mine  pumps  depends  entirely  on  the  type 
of  pump.  When  the  pump  is  of  the  fly-wheel  type,  the  method 
is  that  used  in  Chapter  V.  The  great  length  of  discharge  pipe 
would  mean  considerable  inertia,  but  by  the  use  of  air  chambers 


668 


PUMPING  MACHINERY 


the  discharge  in  the  pipe  is  not  a  fluctuating  one  and  hence 
the  friction  in  this  pipe  line  is  the  only  part  requiring  consider- 
ation in  addition  to  the  static  head.  The  problem  of  the  fly- 
wheel design  is  slightly  different  from  that  considered  in  Chapter 
V  if  an  electric  motor  is  used  to  drive  the  pump.  In  this  case 
the  tangential  effort  of  the  pump  end  pressure,  combined  with 
the  friction  of  stuffing  boxes,  and  forces  from  inertia  and 
weight,  is  found.  From  this  the  excess  and  deficiency  of  tan- 
gential effort  area  is  determined  above  and  below  a  line 


FIG.  500. — Jeanesville  Mine  Pump. 

representing  the  delivered  torque  from  the    motor,  which,  of 
course,  is  constant. 

The  problem  of  a  direct-acting  steam  pump  without  a 
fly  wheel  has  to  be  handled  in  a  different  manner.  The  inertia 
of  the  moving  parts  is  the  important  element  in  this  design. 
The  first  problem  to  be  considered  will  be  one  involving  the 
type  of  mine  pump  shown  in  Fig.  494,  but  in  which  the  fly 
wheel  is  omitted.  In  this  case  it  is  assumed  that  the  indicator 
cards  from  the  two  steam  cylinders  are  those  shown  at  A, 
Fig.  502.  These  are  combined  as  at  B  by  taking  the  height 
to  be 


MINE  PUMPS 


669 


670 


P  UMPING  M  A  CHINER  Y 


Down 


Up 


~~~H|     Resultant  Curve 

W  Water  Curve 
C  Unbalanced  Weight 
/I  Friction 
/2  Friction 
_J  Resultant  Curve 
W2  Water  Curve 


Down 


FIG.  502. — Forces  in  Direct-Acting  Rod  Mine  Pump, 


MINE  PUMPS  671 

This  is  the  force  of  the  steam  acting  on  the  system.  The 
unbalanced  weight  of  the  moving  parts  of  the  whole  system 
per  square  inch  of  low  pressure  piston  area  is  next  found  as 


The  value  of  this  force  which  reaches  the  piston  rod  will 
depend  on  the  inclination  of  the  bell-crank  levers  and  by  con- 
structing the  perpendiculars  from  the  pivot,  ph  and  pv  to  the 
line  of  pump-rod  pull  and  piston-rod  pull,  respectively,  the 
pull  in  the  direction  of  the  piston  rod  may  be  computed. 


This  gives  the  curve  c,  Ci,  c2. 

The  weights  of  all  the  parts  of  the  rods,  balance  weights, 
levers  and  pump  pistons  are  now  found.  These  must  have 
an  acceleration  equal  to  the  acceleration  of  the  piston,  multi- 
plied by  the  ratios  of  the  various  lever  arms  up  to  the  part 
considered.  Thus  the  acceleration  of  the  main  pump  rod  is 


If  T  is  the  ratio  of  the  arms  of  a  balance  lever  the  acceler- 
b 

ation  of  this  part  is 

pva 


If  now  each  weight  is  examined  and  its  acceleration  found 
in  terms  of  ap,  the  force  which  must  be  exerted  on  the  piston 
to  accelerate  all  of  these  parts  is 


672  PUMPING  MACHINERY 


This  means  that  if  the  sum  of  the  terms  7  ^~T  Wx  is  found, 

*~*pk9 

this  will  give  a  weight  which  may  be  assumed  to  have  the  same 
motion  as  the  piston  in  computing  the  force  on  the  piston 

required  to  give  these  parts  their  motion.     These  ratios,  ~, 

Ph 

J-,  etc.,  vary  slightly,  so  that  this  sum  will  change  as  the  piston 

moves.     The  value  of  this  is  laid  off  for  different  positions 
of  the  piston  along  the  stroke,  as  in  C,  Fig.  502. 

The  pressure  per  square  foot  on  the  pistons  from  the  water 
is  equal  to 


if  the  air  chambers  are  of  the  proper  size  to  prevent  the 
fluctuation  of  pressure,  and  the  connections  to  the  air  chambers 
are  so  short  that  the  iriertia  of  the  water  in  this  connection 
can  be  neglected.  If  the  connection  is  not  short,  then  the  weight 
of  the  water  between  the  chamber  and  the  pump  must  be  added 
as  a  weight  of  one  of  the  reciprocating  parts.  In  general,  how- 
ever, the  terms  involving  X  and  a  of  Eq.  (21),  Chapter  V,  are 
so  small  when  compared  with  hi  that  they  may  be  neglected. 

The  sum  of  the  products  whA~j-  for  the  various  pump  pistons 

when  divided  by  AI  will  give  the  pressure  per  square  inch 
which  must  be  exerted  per  square  inch  of  L.P.  piston  required 
to  lift  the  water.  This  is  shown  by  the  curve  w,  wit  w^  etc., 
for  the  different  piston  positions. 

If  there  is  no  air  chamber  on  the  pump,  the  weight  of  the 
whole  column  of  discharging  water  must  be  added  to  the 
weight  of  the  reciprocating  parts  in  the  determination  of  the 
weight  to  be  accelerated. 

The  friction  of  the  various  balancing  and  supporting  levers 
should  next  be  found  and  with  this  the  friction  of  the  rods 
against  their  supports  when  not  vertical  as  well  as  the  friction 
of  the  stuffing  boxes.  These  forces  are  overcome  by  a  force 


MINE   PUMPS  673 

in  the  direction  of  the  piston  travel  which  may  be  found  by 
the  use  of  the  moments  of  the  various  forces  about  different 
pivots.  In  this  way  the  curve  /1?  /2,  fs  .  .  .  is  found,  showing 
the  force  per  square  inch  of  low-pressure  piston  area  which 
must  be  exerted  on  the  piston  in  order  to  overcome  the  friction. 
These  curves  are  combined,  giving  the  resultant  curve  shown 
at  D.  The  point  i  is  found  as 


i  =  all  -alwl  -a 
Ef  =a\bi  —a\Wi  —a 

Since  the  term  a\c\  is  added  in  one  case  and  subtracted  in  the 
other,  it  is  seen  that  the  work  on  the  steam  end  is  less  on  the 
one  stroke  than  on  the  other  unless  the  weights  are  balanced. 
If  under  balanced,  the  work  of  the  steam  end  is  greater  on 
the  up  stroke,  while  if  overbalanced,  the  work  of  the  steam 
end  is  greater  on  the  down  stroke. 

The  area  of  the  curve  /'/i,  which  is  the  resultant  of  the 
pressures  which  act  during  the  stroke,  represents  unbalanced 
work  and  hence  the  positive  and  negative  parts  must  be  equal, 
or  the  moving  parts  of  the  pump  will  not  come  to  rest  at  the 
end  of  the  stroke.  In  the  actual  engine  of  the  direct-acting 
form  it  may  be  said  that  the  stroke  ends,  the  moving  parts 
coming  to  rest,  when  the  negative  area  equals  the  .positive. 
On  the  return  stroke  the  same  must  be  true. 

If  now  the  net  unbalanced  pressure  on  each  square  inch 
of  piston  area  at  some  point  of  the  stroke  be  divided  by  the 
equivalent  weight  per  square  inch  of  piston  area  at  that  point 
and  multiplied  by  g,  the  result  will  be  the  acceleration  in  the 
parts  caused  by  the  force,  or 

ap=%7 

If  now  a  curve  be  plotted  with  the  values  of  ap  for  different 
points,  the  result  will  show  how  the  acceleration  varies. 
This  is  shown  in  Fig.  503.     Now 

dv     dv     ds        dv 


674 


PUMPING  MACHINERY 


Hence 


and 


-y 
ds 


;r       V2 
Vdv=\ads=—. 


TIP    b 


Down 


FIG.  503. — Acceleration  and  Velocity  Curves. 

The  area  of  the  curve  just  drawn  from  the  end  to  any  point 

V2 
is  — .     If  then  the  integral  curve  be  drawn  in  Fig.  503  the 

V2 
ordinates  of  this  curve  will  be  —  and  from  these  V  may  be 

found,  giving  the  dotted  curve.  From  this  curve  the  velocity 
time  curve  may  be  constructed  as  shown  in  Fig.  193  and  from 
it,  the  time  taken  to  complete  one  stroke.  The  curve  is  con- 
structed for  the  two  strokes.  The  time  found  will  determine 
the  number  of  strokes  per  m'nute  and  consequently  the  capacity 
of  the  pump. 


MINE  PUMPS  675 

If  it  is  desired  to  cut  down  the  speed  of  the  pump  more, 
a  balanced  mass  could  be  added  to  the  system,  causing  the 
acceleration  curve  to  be  lower,  although  the  force  diagram, 
Fig.  502,  would  not  change.  This  may  also  be  accomplished 
by  decreasing  the  maximum  steam  pressure  and  carrying  the 
cut  off  later.  The  total  area  of  the  indicator  card  cannot  be 
changed,  as  that  must  be  equal  to  the  work  done  on  the  water 
and  in  overcoming  the  friction. 

If  it  is  desired  to  increase  the  speed  of  the  pump,  the 
balanced  mass  must  be  made  smaller  if  possible  and  if  this 
cannot  be  done  the  steam  pressure  is  increased  and  the  cut  off 
is  made  earlier.  The  first  of  these  changes  the  acceleration 
curve,  increasing  its  height  without  any  change  in  the  curve 
of  unbalanced  force,  while  the  second  method  increases  the 
height  of  the  acceleration  curve  by  increasing  the  height  of 
the  curve  of  unbalanced  force,  since  there  is  no  change  in  the 
weight  of  the  moving  parts  in  this  case.  In  the  latter  case 
also,  it  must  be  remembered  that  the  area  of  the  indicator 
cards  cannot  be  changed. 

The  mass  in  some  cases  may  be  so  great  that  even  with  high 
steam  pressure  the  speed  must  be  slow  and  the  capacity  small. 
In  such  cases  balance  weights  may  be  omitted;  then  it 
becomes  necessary  to  cut  down  the  amount  of  steam  used  on 
the  down  stroke.  A  better  method  is  that  shown  in  Fig.  495, 
in  which  two  rods  and  pump  sets  are  so  used  that  as  one  is 
descending  the  other  is  ascending.  In  this  case  the  effect  is 
a  balanced  one  without  the  inertia  effect  of  balance  weights, 
for  each  side  of  the  apparatus  is  designed  to  do  pumping,  and 
the  mass  is  that  of  an  unbalanced  pump.  Such  an  arrange- 
ment will  give  the  fastest  moving  pump  of  this  type. 

In  using  these  pumps  it  is  well  to  note  that  should  the  steam 
pressure  be  increased  without  the  reduction  of  the  cut  off, 
the  positive  area  of  unbalanced  force  would  be  greater  than 
the  negative  area,  and  consequently  the  piston  would  strike 
the  head  with  considerable  force.  For  this  reason  a  large 
cushion  space  should  be  used  or  springs  should  be  applied 
to  take  up  the  shock.  In  case  the  pressure  is  reduced  the 


676 


PUMPING  MACHINERY 


piston  will  come  to  rest  as  soon  as  the  negative  area  is  equal 
to  the  positive  area.  In  some  cases  the  steam  may  be  admitted 
to  the  opposite  end,  at  which  time  the  full  boiler  pressure  will 
act  to  stop  the  motion  and  reverse  the  pump.  This  would 
mean  a  large  increase  in  the  negative  residual  pressure,  which 
would  mean  a  rapid  negative  acceleration,  bringing  the  parts 
to  rest. 

The  use  of  direct-acting  pumps  with  steam  cylinders  at  the 
pump  involves  the  same  principles  of  design  as  the  pumps 
just  considered.  In  this  case,  however,  the  total  mass  recip- 
rocated is  quite  small  and  for  that  reason  the  residual  steam 

Combined  H.P.  and  L.P.  Steam 


Water  Resistance 


FIG.  504.  —  Compound  Direct-acting  Pump    Diagram. 

pressure  cannot  be  as  large  as  used  with  the  other  type   of 
pump,  since  the  acceleration 


would  be  very  great.  This  would  cause  pounding  and  there 
would  be  some  danger  of  wrecking  the  pump  if  the  valve  gear 
did  not  work  properly.  For  this  reason  the  method  is  to 
carry  the  pressure  practically  the  full  length  of  the  stroke,  get- 
ting the  advantages  of  expansion  in  the  use  of  several  cylinders 
in  which  the  cut  off  is  at  or  near  the  end  of  the  stroke.  For 
such  the  combined  card  is  shown  in  Fig.  504.  The  steam 
pressure  is  slightly  above  the  resistance  for  the  major  part  of 
the  stroke  and  the  piston  is  brought  to  rest  by  the  cushion 
steam  or  the  valve  reversal. 


MINE  PUMPS  677 

The  principles  of  Chapter  V  may  be  applied  in  this  case 
for  the  resistance  of  the  various  parts,  the  terms  which  depend 
on  the  acceleration  being  so  small  as  to  be  neglected  in  com- 
parison with  the  height  through  which  the  water  is  lifted. 
Should  this  be  such  that  these  terms  are  not  small,  they  would 
have  to  be  neglected  in  the  first  approximation,  when  the 
motion  is  found;  then  computed  for  the  motion  thus  deter- 
mined, and  a  second  approximation  made  in  which  they  are 
used.  If  these  terms  are  changed  much  by  the  new  curve 
of  time  and  velocity  a  third  approximation  should  be  made. 

In  this  manner  the  probable  capacity  of  the  pump  may  be 
found.  Of  course  in  this  type  of  pump  with  a  proper  cushion 
chamber  the  speed  may  be  increased  by  increasing  the  steam 
pressure.  After  the  speed  reaches  its  limit,  pounding  begins, 
as  the  cushion  cannot  care  for  the  energy  stored  up  in  the 
moving  parts. 


BIBLIOGRAPHY 


THE  list  of  references  given  below  have  been  used  in  the  preparation  of 
this  treatise.  The  references  are  given  by  volume  number  or  date  with  the 
page  on  which  the  reference  is  found.  The  following  abbreviations  are  used: 

Am.  Mch.  American  Machinist. 

Engng.  London  Engineering. 

Eng.  Mag.  Engineering  Magazine. 

Eng.  News  Engineering  News. 

Eng.  Rec.  Engineering  Record. 

Engr.  The  Engineer  (American). 

A.S.C.E.  Transactions  of  the  American  Society  of  Civil  Engineers. 

A.S.M.E.  Transactions  of  the  American  Society  of  Mechanical  Engi- 
neers. 

J.A.S.M.E.  Journal  of  the  American  Society  of  Mechanical  Engineers. 

N.E.W.W.  Journal  of  the  New  England  Water  Works  Association. 

A.W.W.A.  Transactions  of  the  American  Water  Works  Association. 

Z.  f.  d.  g.  Turb.  Zeitschrift  fiir  das  gesampte  Turbinenwesen. 

Z.  d.  V.  d.  Ing.  Zeitschrift  des  Vereines  deutscher  Ingenieure. 

Elec.  World  Electrical  World. 

Power  Power  and  the  Engineer. 

A.E.S.  Journal  of  Associated  Engineering  Societies. 

Cassiers  Cassiers  Magazine. 

These  are  arranged  under  the  following  heads: 


Texts. 

Air  Pump  and  Screw  Pump  Tests. 
Centrifugal  Pumps,  Theory  and  De- 
sign. 

Centrifugal  Pumps. 
Centrifugal  Pump  Stations. 
Centrifugal  Pump  Tests. 
Costs. 

Fire  Pumps  and  Stations. 
Historical. 
Injectors  and  Direct  Steam  Pumps. 


Irrigation,  Drainage  and  Sewage  Pumps 
and  Stations. 

Mine  Pumps. 

Reciprocating  Pumps,  Tests  and  Re- 
sults. 

Reciprocating  Pumps,  Design  and 
Theory. 

Rotary  and  Disc  Pumps. 

Simplex  Pumps. 

Special  Pumps. 

Water  Works  and  Water  Works  Pumps. 

679 


680  BIBLIOGRAPHY 


TEXTS 

Centrifugal  Pumps  and  Turbines.  Chas.  H.  Innes. 

A  Descriptive  and  Historical  Account  of  Hydrau- 
lic and  Other  Machines  for  Raising  Water.  Thomas  Ewbank. 

Encyclopedia  Britannica.  Article  "Steam  Engine." 

Growth  of  Steam  Engine.  R.  H.  Thurston. 

Hydraulic  Motors.  G.  R.  Bodmer. 

Hydraulic  Power  Engineering.  G.  Croydon  Marks. 

Hydraulics.  Mansfield  Merriman. 

Hydraulics.  F.  C.  Lea. 

Hydraulics.  A.  H.  Gibson. 

Hydraulics.  W.  C.  Unwin. 

Irrigation  Works  in  India.  R.  B.  Buckley. 

Lives  of  Boulton  and  Watt.  Smiles. 

Life  of  Newcomen.  Smiles. 

Lives  of  George  and  Robert  Stevenson.  Smiles. 

Machine  Design.  W.  C.  Unwin. 

Machine  Drawing  and  Design.  Low  and  Bevis. 

Mechanical  Engineers'  Pocket  Book.  Kent. 

Die  Pumpen.  Hartmann-Knoke-Berg. 

Practice  and  Theory  of  the  Injector.  Strickland  Kneass. 

Pumping  Machinery.  Wm.  M.  Barr. 

Water  Supply  and  Irrigation  Papers  of  the  U.  S. 

Geological  Survey  No.  i.  Herbert  M.  Wilson. 

Water  Works  Pumps.  Chas.  A.  Hague. 

Die  Zentrifugal  Pumpen.  Fritz  Neumann. 

AIR  PUMP  AND  SCREW  PUMP  TESTS 

Experiments  on  Air  Pumps.  Engng.,  86:  703. 

Screw  Pumps.  Eng.  Rec.,  58:    426;    Eng. 

News,  60:  269. 

Air  Lift  Pumps  at  Redlands,  Cal.  Eng.  Rec.,  51:  8. 

Air  Lift  Pump.  Eng.  News,  32:  27;  Engr., 

Aug.  15,  1904. 

Merrill  Compressed  Air  Pump.  Engr.,  July0! 5,  1904. 

Air  Lift  Pump.  Engr.,  June  i,  1906. 

CENTRIFUGAL  PUMPS,  THEORY  AND  DESIGN 

On  the  Construction  of  the  Impellers  of  High- 
Speed  Low-head  Centrifugal  Pumps,  by  P.  Zeit.  f.  d.  g.  Turb.,  Nov.  10, 
Riebensahm.  1909. 

Investigation  of  Centrifugal  Pumps,  Part  II.     C. 

B.  Stewart.  Bui.  U.  of  Wis.,  No.  318. 

Design  of  High  Lift  Centrifugal  Pumps,  by  F.  z. 

Nedden.  Eng.  Mag.,  Jan.-May,  1910. 


BIBLIOGRAPHY  681 

Le    Calcul   de   la   hauteur  de    refoulement    des  Le    Genie    Civil,    Nov.     6, 

pompes  centrifuges.  1909. 

Notes    on    Centrifugal    Pump    Design.      J.    B. 

Sperry.  Am.  Mach.,  Nov.  19,  1908. 

Die  Grundlagen  der   Lorenzschen  Theorie  des  Zeit.  f.  d.  g.  Turb.,  May  10, 

Kreiselrader.     R.  Lowy.  1909. 

Zur  Theorie  der  Zentrifugal  pumpen.     E.  Busse.  Zeit.  f.  d.  g.  Turb.,  Jan.  9- 

Feb.  10,  1909. 
Ausfiihrungen  und  Versuchsresultate  von  Hoch-  Zeit.  f.  d.  g.  Turb.,  Feb.  20, 

druckzentrifugalpumpen,  by  Griessman.  1909. 

Die    Wirkungsweise    der     Kreiselpumpen     und  Mit.    u.     Forsh.,    Oct.    12, 
Ventilatoren.     Dr.  R.  Biel.  1907,  Eng.  Rec.,  Oct.  12, 

1907. 
The  Design  of  Centrifugal  Pumps,  by  J.  Richards.  Eng.    News,  July    29    and 

Aug.  5,  1897. 
Design  of  Centrifugal  Pumps,  E.  U.  Percy.  Jour.  Elec.  Power  and  Gas, 

April  27,  1907. 
Theory  of  Centrifugal  Pumps  and  Fans.  A.S.C.E.,  50,  No.  963;   51, 

No.  936. 

Design  of  a -Centrifugal  Pump.  Am.  Mach.,  Sept.  22,  1910. 

Ueber  der  hydraulischen  Wirkungsgrad  von  Tur- 
binen  bei  ihrer  Verwendung  als  Kraftma- 
schinen  und  Pumpen.  R.  Proell  (Springer,  1904). 


CENTRIFUGAL  PUMPS 

Allis-Chalmers  Electrically  Driven  Fire  Service,  Power,  Nov.  2,  1909,   760; 

New  York,  79%  Eff.    '  J.A.S.M.K,  Sept.,  1909. 

Worthington,  Electrically  Driven,  80%  Eff.,  94.2  Engng.,  Feb.  5,  1909;  Zeit. 

M.  head.  f.  d.  g.  Turb.,  Sept.  20, 

1909. 

A  New  Turbine  Pump.  Engr.,  June  14,  1907. 

Some    Practical    Experiences    with    Centrifugal 

Pumps  for  Water  Works  Service,  C.  A.  Hague  Eng.  News,  Aug.  i,  1907. 
Tan -gyro  Centrifugal  Pump.  Engng.,  Jan.  6,  1911. 

The  Worthington  Turbine  Pump.  Am.  Mach.,  27:  514. 

Fire  Boats  of  Chicago.  Fire  &  Water  Eng.,  April  i, 

1908. 

Large  Centrifugals.  Engng.,  42:  233;  38:  12. 

Appold  Pumps  at  Portsmouth  Docks.  Engng.,  2:  382. 

Stadil  Fjord  Pumps.  Engng.,  i:  13. 

Andrews  Pumps.  Engng.,  3:  495. 

Gwynne  Pumps.  Engng.,  i:  82;  6:  185;  50: 

755- 
High  Lift  Centrifugals.  Engng.,  6:  519;  Eng.  Rec., 

58:  112. 
Bernays  Pump.  Engng.,  15:  456. 


682  BIBLIOGRAPHY 

New  York  Dredging  Pumps.  A.S.C.E.,  25:  599. 

Irrigation  Pumps.  A.S.C.E.,  54:  163. 

Dredging  Pumps.  A.S.C.E.,    54:     285;     54: 

391;  A.E.S.,  19:  140; 

Eng.  News,  27:  290. 

Vertical  Shaft  Centrifugals.  Engng.,  37:  138. 

Malta  Dock  Pumps.  Engng.,  44:  224. 

Southwark  Foundry  Pumps.  Eng.  News,  14:  349. 

Compound  Centrifugals.  Eng.  News,  26:  246. 

Balancing  Centrifugals.  Eng.  News,  25:  114. 

Lea-Degen  Pumps.  Eng.  Rec.,  54:  352. 

Well  Pumps.  Eng.  Rec.,  50:  177. 

High  Lift  Centrifugals.  Eng.  Rec.,  58:  112. 

Helicoidal  Pumps.  Engng.,  42:  570. 

Twin  Pumps  for  Dock  Service.  Engng.,  86:  719. 


CENTRIFUGAL  PUMPS'   STATIONS 

New  York  Fire  Station.  Power,  Nov.  2,  1909;  A.S. 

M.E.,   31:    437;    Eng. 

Rec.,  57:  22. 

Auxiliary  Pumping  Station  at  Charleston,  W.  Va.  Eng.  Rec.,  Sept.  25,  1909. 
The  Turbine  Pumps  of  Montreal  Water  Works.    Eng.  Rec.,  54:  488. 
Chicago  Fire  Boats.  Fire  &  Water  Eng.,  April  i, 

1908. 

Electric  Mine  Drainage  in  Europe.  Elec.  World,  Nov.  17,  1906. 

Electric   Pumping   Equipment   for   the   Mexico  Eng.  Rec.,  Dec.  8,  1906,  58: 

Water  Works.  128. 

Low  Head  Pumping  Plant  at  New  Orleans.  Eng.  Rec.,  April  23,  1910. 

Peoria  Station.  Eng.  Rec.,  51:  139. 

Montreal  Water  Works.  Eng.  Rec.,  54:  488. 

Schenectady  Station.  Eng.  Rec.,  51:  640. 


CENTRIFUGAL  PUMPS— TESTING 

Tests  of  a  New  Centrifugal  Pump  (78%  EfL).        Eng.  Rec.,  54:  352. 
Versuche  an  einer  Zentrifugalpumpe.  Z.  f.  d.  g.  Turb.,  March  10- 

20,  1908. 

Twin  Centrifugals  for  Dock  Service.  Engng.,  86:  719. 

Centrifugal  Pumps  of  1885.  Engng.,  40:  124,  215. 

Early  Centrifugals.  Engng.,  2:  382. 

Centrifugal  Pump  Efficiency.  A.S.M.E.,  7:  598. 

Centrifugal  Pump  Test.          ,  Engng.,  52:  696. 

Allis-Chalmers  Pumps  at  Pittsburg.  Eng.  Rec.,  58:    649;   Eng. 

News,  60:  573. 
Richards  Tests.  Eng.  News,  38:  75. 


BIBLIOGRAPHY 


683 


De  Laval  Centrifugal. 
Worthington  Turbine  Pumps. 

Sewage  Pumps  at  Carlisle. 
Vertical  Shaft  Pump. 
Drainage  Pumps  of  the  South. 
Pacific  Coast  Pumps. 
Duty  Test  at  Torresdal  Station. 


Eng.  Rec.,  50:  216. 
Engng.,    87:      181; 

News,  58:  112. 
Engng.,  87:  46. 
Engng.,  37:  138. 
A.S.C.E.,  56:  159. 
A.S.C.E.,  56:  148. 
Power,  Nov.  16,  1909. 


Eng. 


COSTS 

Cost  of  2o-Million  Gallon  Pumps. 

Cost  of  Large  Pumps  in  1870. 

Cost  of  Pumps  in  1886. 

Cost  of  Pumps. 

Cost  of  Pumps  in  1886. 

Cost  of  Piping  in  1884. 

Cost  of  Holly  Engines. 

Cost  of  Large  Pumps. 

Cost  of  Pumps  for  Chicago. 

Cost  of  Small  Plants. 

Comstock  Lode  Pumps. 

Cost  of  Boston  Plants. 

Cost  of  Raising  Water. 

Cost  of  Pumping. 

Cost  of  Raising  Water. 

Cost  of  Pumping  Water. 

Cost  of  Electric  Pumping. 

Cost  of  Operating  Producer  Gas  Stations. 

Oil  Engine  Stations. 

Olean  Gas  Engine  Station. 

Commercial  Pumping  Engine. 

Steam  Driven  vs.  Electric  Driven  Pumps. 

High  vs.  Low  Duty  Pumps. 

Possibilities  of  Economy  in  Pumping  Engines. 

Small  Pumping  Plants. 

Duty  and  Costs  of  Operating  Pumps. 

Water  Waste  in  New  York. 

Receipts  from  W.  W.  of  the  Eight  Largest  Cities  in 

America,  for  1884. 

Financial  Management  of  Water  Works. 
Depreciation. 

American  Water  Works  Statistics. 
Comparative  Merits  of  Various  Pumps. 


Eng.  News,  23:  578. 
Eng.  News,  15:  93. 
Eng.  News,  16:  93,  95. 
Eng.  News,  13:  340. 
A.E.S.,  14:  24. 
Eng.  News,  12:  47. 
Eng.  News,  30:  67. 
Eng.  News,  30:  342, 

435,  484- 

Eng.  News,  32:  225. 
N.E.W.W.,  14:  163. 
Eng.  Rec.,  51:  360. 
N.E.W.W.,  15:  299. 
Van  Nos.  i :  648. 
A.S.C.E.,  4:  369. 
Eng.  News,  16:  231. 
Eng.  News,  30:  181. 
N.E.W.W.,  10:  184. 
Eng.  Rec.,  59:  786. 
Eng.  Rec.,  58:  230. 
Eng.  Rec.,  58:  112. 
Eng.  Rec.,  51:  537. 
A.W.W.A.,  1907,  189. 
A.W.W.A.,  1908,  725. 
N.E.W.W.,  13:  163. 
A.W.W.A,  1907,  157. 
A.W.W.A.,  1907,  210. 
Eng.  Rec.,  48:  340. 

Eng.  News,  12:  47. 
N.E.W.W.,  ii :  63. 
A.W.W.A,  1903,  473. 
Eng.  News,  13:  340. 
Engng.,  37:  394. 


424, 


684 


BIBLIOGRAPHY 


FIRE  PUMPS  AND   STATIONS 


Arrangement  of  Hydrants  and  Water  Pipes. 
Capacity  of  Steam  Fire  Engines,  Hydrants  and 

Hose. 

Fire  Protection,  Amount  of. 
High  Pressure  Stations. 

Boston. 

Chicago. 

New  York. 

Philadelphia. 

Providence. 

Hand  Pump,  Horse-drawn. 
Gould  System. 
Building  a  Fire  Pump. 
Merryweather  Pump. 


N.E.W.W.,  7:  49. 

N.E.W.W.,  9:  151. 
N.E.W.W,  3:  97. 
N.E.W.W.,  13:  304. 
Eng.  Rec.,  48:  138. 
Eng.  Rec.,  57:    22;   J.A.S. 

M.E.,  Sept.,  1909. 
A.S.M.E.,3i:  437;  Power, 

1909,  760. 

Eng.  Rec.,  59:  748;  49:  309. 
N.E.W.W.,  13:  85. 
Engng.,  15:  301. 
Engng.,  21 :  432. 
Am.  Mach.,  Jan.  5,  1911. 
Engng.,  10;  192, 


HISTORICAL 

See  Text-Book  List. 

Centrifugal  Pumps. 

Early  History  of  Centrifugal  Pump. 

Development  of  Centrifugal  Pumps. 

Invention  of  Centrifugal  Pump. 

History  of  Centrifugal  Pump. 

Engineering  300  Years  Ago. 

Fire  Pumps. 

Flash  Pumps  in  Holland. 

Humphrey  Pump. 

Hydraulic  Rams. 

Newcomen  and  His  Work. 

Pumping  Machinery. 

Pumping  Machinery. 

Development  of  Pumping  Machinery. 
Some  Heavy  Modern  Pumping  Machinery. 
Worthington,  Henry  R. 


Cassiers,  28:  154. 
Engng.,  i:  275. 
A.W.W.A.,  1904,  175. 
Engng.,  50:  670. 
Prac.  Mech.  Jqur.,  1851. 
Cassiers,  8:  97. 
Cassiers,  7:  307. 
Eng.  News,  May  19,  1910. 
Engng.,  88:  737,  512,  514. 
Cassiers,  28:  65. 
Cassiers,  Dec.,  1891. 
Engr.,  Nov.  16,  1903. 
.Eng.  Mag.,  May,   1891,  i: 

141. 

Eng.  Mag.,  5:  451. 
Cassiers,  Jan.,  1895. 
Eng.  News,  Sup.,  March  23, 

1893. 


INJECTORS  AND   DIRECT   STEAM  PUMPS 


American  Steam  Jet  Pump. 
Direct  Displacement  Pump. 
Korting  Bilge  Pump. 
Pulsometer. 


Engng.,  2:  237. 
Engng.,  March  15,  1905. 
Engng.,  19:  477. 
Engng.,  July  15,  1905. 


BIBLIOGRAPHY 


685 


Pulsometer,  Hall's. 
Pulsometer,  Test. 
Steam  Ejector. 
Water  Jet  Pump. 


Engng.,  22:  56. 
Tech.  Quar.,  Sept.,  1901. 
Engng.,  ii :  416. 
Engng.,  i:  117. 


IRRIGATION,   DRAINAGE,   AND   SEWAGE   PUMPS  AND 
STATIONS 


Automatic  Sewage  Station. 
Boston  Sewage  Engine  and  Plant. 
Carlisle  (Eng.)  Plant. 
Centrifugals  for  Irrigation. 
Cheswick  Drainage  Pump. 
Chicago  Sewer  System. 


Dorchester  Leavitt  Pump. 

Dortdrecht  Drainage  Station. 

Ferrara  Drainage  Station. 

Flash  Pumps  in  Holland. 

Hampton  Institute. 

Hull  Plant. 

Huntly  Plant. 

Irrigation  on  the  Pacific  Coast. 

Leavitt  Pumps. 

Memphis  Storm  Sewer  Station. 

New  Orleans  System. 

Portsmouth  Pump. 

Providence  Pumps. 

Scoop  Wheels. 

Screw  Pump  for  Milwaukee. 
Stadil  Fjord  Pump. 
Thompson  Sewage  Valves. 
Triplex  Drainage  Pump. 
Washington,  D.  C,  Plant. 
McLaren  Sewage  Pump. 


Eng.  Rec.,  57:  272. 
Engng.,  40:  555. 
Engng.,  87:  46. 
A.S.C.E.,  54:  163. 
Engng.,  27:  477. 
Eng.   Rec.,   52:    578;    58: 

426;     Eng.    News,    60: 

269. 

Eng.  Rec.,  51:  676. 
Engng.,  21 :  192. 
Engng.,  21 :  9. 
Eng.  News,  May  19,  1910. 
Eng.  Rec.,  52:  566. 
Eng.  News,  23:  126. 
Eng.  News,  60:  262. 
Eng.,  44:  479. 
N.E.W.W.,  9:  163. 
Eng.  Rec.,  53:  496. 
Eng.  Mag.,  25:  342. 
Engng.,  10:  44. 
Eng.  News,  32 :  259. 
Engng.,  8:  174;  9:  183, 

194,  230,  274,  321,  441. 
Eng.  News,  23:  218. 
Engng.,  i:  13. 
Engng.,  3:  127. 
Engng.,  29:  23. 
Eng.  Rec.,  58:  200. 
Engng.,  19:  317. 


MINE  PUMPS 


Allis  Pumps  at  Chapin  Mine. 
Bohemia  Pumps. 
Cameron  Pump. 
Comstock  Lode  Pumps. 

Cornish  Pump  at  Ontario  Mine. 

Deep  Mine  Pumps. 

Duplex  Sinking  Pump,  Electric  Drive. 


Eng.  News,  30:  310. 
Engng.,  27:  155. 
Engng.,  14:  180. 
Eng.  Rec.,  51:  360;  Engr., 

May  15,  1905. 
Eng.  News,  32:  440. 
52  :   40. 


Eng.  News,  29:  471. 


BIBLIOGRAPHY 


Electric  Drive. 

Electric  Driven  Knowles  Pump. 
Eston,  Eng.  Pumps. 

Fowey  Consols  Mines  (Recent  Practice  in  Pump- 
ing). 
High  Speed  Pumps. 

Mine  Pumps  by  Howell  Green. 

Mine  Pumping. 

Mine  Working  at  Great  Distances  with  Rods. 

Rittinger  Pump. 

Underground  Mine  Pump. 


Engng.,  44:  534. 
Eng.  Rec.,  54:  403. 
Engng.,  16:  294. 

N.E.W.W.,  8:  85. 
Eng.  Mag.,  24:  772: 

July  15,  1904. 
A.S.M.E.,  4:  217. 
Cassiers,  31:  125. 

Engng-,  3i:  281. 
Engng.,  30:  151. 
Engng.,  43:  285. 


Engr, 


RECIPROCATING   PUMPS— TESTS  AND   RESULTS 


Allegheny  Allis  Pump. 

Belfast  Engine. 

Belmont  Pump. 

Blake  Pump. 

Boston  Allis  Pump. 

Brooklyn  Pumps. 

Brooklyn  Pump. 

Chestnut  Hill  Pump. 

Chicago  Allis  Pump. 

Cleveland  Snow  Pump. 

Coal  Consumption  of  Pumps  in  1867. 

Cornish  Mine  Pump. 

Davison  Pump  at  Norwood. 

Davy  Pump. 

Duties  on  Cornish  Engines. 

Fall  River  Davison  Pump. 

Farcot  Pump. 

Gaskill  Pump. 

Hannibal  Pump. 

Hathorn,  Davey  &  Co. 

High  Duty  Tests. 

Improvements  in  Pumping  Engines. 

Lawrence  Pump. 

Lowell  Pump. 

Memphis  Worthington  Pump. 

Moreland  Pump. 

Newark  Pump. 
Norwood  Pump. 
Odessa  English  Pump. 
Pawtucket  Pump. 


Engng.,  41:  33. 
Eng.  Rec.,  51:  517. 
Engng.,  15:  411. 
Eng.  News,  29:  137. 
N.E.W.W.,  June,  1901. 
Eng.  News,  28:  561. 
Engng.,  9:  369. 
Eng.  News,  28:  578. 
Eng.  News,  30:  149. 
Eng.  Rec.,  48:  341. 
Engng.,  4:  122. 
Eng.  News,  32:  440. 
Eng.  News,  15:  254. 
Engng.,  22:  421. 
Eng.  News,  22:    602; 

Engng.,  i:  107. 
Eng.  News,  10:  423. 
Engng.,  26:  70. 
Eng.  News,  n:  118.   . 
Eng.  News,  14:  142. 
Engng.,  86:  37. 
Eng.  Mag.,  20:  281. 
N.E.W.W.,  March,  1899. 
Engng.,  27:  58. 
Eng.  News,  27:  374. 
Eng.  News,  22:    151;    26: 

233- 
Engng.,  38:   319,  384,  472, 

520. 

Engng.,  10:  224. 
Eng.  News,  15:  254. 
Eng.  Rec.,  50:  638. 
Engng.,  28:  189. 


BIBLIOGRAPHY 


687 


Pontiac  Pump. 

Recent  Practice. 

Richardson  Triple  Expansion  Pump. 

Superheated  Steam  Test. 

Trenton  Allis  Pump. 

Water  Works  Pumps  of  1875. 

Worthington  Pumps. 


Eng.  Rec.,  48:  280. 
Eng.  News,  30:  119. 
Engng.,  50:  158. 
Eng.  Rec.,  59:  788. 
Eng.  News,  32:  483. 

A.S.C.E.,  Vol.  4. 

Engng.  News,  16:  289;  30: 

230;  27:   167;    Engng., 

42:    340. 


RECIPROCATING   PUMP   DESIGN   AND   THEORETICAL 
ARTICLES 


Pump  Parts. 

Steel  Forgings. 

Specifications  for  St.  Louis. 

Pump  Piping. 

Foundations  for  Pumps. 

Reducing  Water  Ram  in  Direct  Acting  Engine. 

Steel  Pipe. 

Gutermuth  Valves. 

Ueber  Freigehende  Pumpenventile. 

Valves  on  Hydraulic  Pumps. 

Packing  for  Hydraulic  Pressure. 

Experimental  Study  of  the  Resistance  of  Flow  of 

Water  in  Pipes,  Saph  and  Schoder. 
Superheated  Steam  with  Pumps. 
Friction  in  Pumping  Mains. 
Flow  of  Water  through  Pipes. 
Test  of  Air  Lift  Valves. 
Flow  of  Water  in  Pipes,  by  Williams,  Hubbell, 

and  Fenkell. 


A.W.W.A.,  1005,  151. 
N.E.W.W.,  12:  120. 
N.E.W.W.,  ii :  172. 
Engr.,  April  i,  1905. 
Cassiers,  31:  42. 
N.E.W.W.,  15:  493- 
N.E.W.W.,  13:  314. 
Engng.,  79:  391. 
Z.  d.  V.  d.  Ing.,  March  25, 

1905. 

Am.  Mach.,  April  14,  1910. 
Am.  Mach.,  Sept.  22,  1910. 

A.S.C.E.,  54:  253. 
A.S.M.E.,  21. 
N.E.W.W.,  10:  234. 
Cassiers,  29:  22. 
N.E.W.W.,  u:  51. 

A.S.C.E.,  Vol.  47. 


ROTARY  AND   DISC   PUMPS 


Behrens  Rotary  Pump. 
Bennison  Rotary  Pump. 
Boulton  and  Imray  Helical  Pump. 
Mauley  Rotary  Pump. 
McFarland  Rotary  Pump. 
Oscillating  Pump. 
Clark  Patent  Rotary  Pump. 
Phillips  Rotary  Pump. 
Portland  Rotary  Pump. 
Stannah  Pendulum  Pump. 
Von  Mottoni  Pendulum  Pump. 


Engng., 
Engng. , 
Engng., 
Engng., 
Engng. , 
Engng. , 
Engng., 
Engng., 
Engng., 
Engng., 
Engng., 


10 :  200. 
19:  69. 
14:  196. 

25:  45i. 

20:  332. 

18:  262. 

44:  187. 

39:  351- 

33*  59- 

23:  56. 

10 :  485. 


688 


BIBLIOGRAPHY 


SIMPLEX  PUMPS 


Blake  Pump. 

Bradfer  Pump. 
Baummann  Pump. 
Cameron  Pump. 

Cherry  Pump. 

Clarkson  Pump. 

Cope  &  Maxwell  Pump. 

Croyden  Water  Works  Pump. 
Davey  Differential  Valve. 
Davison  Pump. 

Deane  Sinking  Pump. 

Decker  Bros.  Pump. 

Deep  Well  Pump. 

Direct  Acting  Simplex  Mine  Pump. 

Earle  Pump. 

Hall  Pump. 

Imperial  Pump. 

Parker  and  Weston  Pump. 

Plam  &  Co.  Pump. 

Pickering  Pump. 

Ramsbottom  Pump. 

Shanks  Pump. 

Silver  Pump. 

Stone  Pump. 

Walker  Pump. 

Walker  and  Holt  Pump. 


Engng.,  20:  37;  Eng.  News, 

7:  91. 

Engng.,  Feb.   19,   1904. 
Engng.,  9:   293;  12:   237. 
Engng.,  5:    253:    4-    213; 

14:  180. 
Engng.,  22:  77. 
Engng.,  18:  210. 
Engng.,   7:    334;    ii :    46; 

22:  57;  35-  327- 
Engng.,  24:  356. 
Engng.,  19:  273;  26:  197. 
Eng.,   News,   8:    436;   12: 

220. 

Engng.,  36:  125. 
Engng.,  16:  371. 

Engng-,  5°:  528. 
Engng.,  14:  180. 
Engng.,  3:  625. 
Eng.  News,  14:  247. 
Engng.,  24:  29,  36. 
Engng.,  22:  121:  24:  183. 
Engng.,  39:  587. 
Engng.,  20:  366. 
Engng.,    9:  192. 
Engng.,  32:  35. 
Engng.,  21 :  268. 
Engng.,  30:  138. 
Engng.,  20:  44. 
Engng.,  ii :  100. 


SPECIAL  PUMPS 


Air  Lift  Pumps. 

Air  Lift  Pump  at  High  Bridge. 

Air  Lift  Pump. 

Bernay's  Steam  Pump. 

Dock  Pump. 

Duplex  Pump  with  Single  Valve. 

High  Pressure  Gas  Power  Station. 

Humphrey's  Gas  Operated  Pump. 


Eng.  News,  29:  541;  32:  27. 
Eng.  Rec.,  49:  672;  A.S.C. 

E.,  54:  i. 

Eng.  News,  50:  675. 
Engng.,  39^  525- 
Engng.,  Feb.  26,  1904. 
Engng.,  42:  168. 
Eng.,  Oct.  i,  1906. 
Engng.,  1909,  737;  1909, 

512-14;  Am.  Mach., 

1911,  Jan.  5. 


BIBLIOGRAPHY 


Hydraulic  Pressure  Pump. 

Hydraulic  Engine  Pump  for  New  London. 

Hydraulic  Pump  Machines. 

Hydraulic  Pumping  Plant  at  Gloucester,  Eng. 

Hydraulic  Ram,  Large. 

Hydraulic  Rams. 

Natural  Gas  Pump. 

Noria  at  Hannah. 

Oil  Pumping  Station. 

Oil  Pumps,  Worthington. 

Oscillating  Pump  (Undulating). 

Oscillating  Pump. 

Producer  Gas  Pumping  Plant. 

Application  of  Gas,  Gasoline,  and  Oil  to  Pumps. 

Rubber  Pump,  Hard. 

Smith's  Explosion  Pump. 

Steam  Turbines  for  Water  Works. 

Valves  for  High  Speed  Pumps. 

Well  Pump  without  Suction  Valves. 

Theory  of  Air  Lift  Pump,  E.G.  Harris. 


Am. 


Eng.  News,  31:  278. 
Eng.  News,  29:  65. 
N.E.W.W.,  i:  34. 
Engng.,  Feb.  12,  1904. 
A.S.C.E.,  54:  159- 
Cassiers,     28:      65; 

Mach.,  April  14,  1910. 
Eng.  Rec.,  50:  712. 
Eng.  News,  19:  159. 
Eng.  Rec.,  57:  676. 
Engng.,  40:  108. 
Eng.  News,  30:  70. 
Engng.,  18:  262. 
A.W.W.A.,  1908,  61. 
N.E.W.W.,  13:  206:    Fov 

er,  Dec.,  1903. 
Eng.  News,  28:  477. 
Eng.  News,  May  19,  1910. 
A.W.W.A.,  1905,  302. 
Engng.,  87:  662. 
Engng.,  40:  55. 
A.S.C.E.,  54:  i. 


WATER  WORKS  AND  WATER  WORKS  PUMPS 


Allegheny  Leavitt  Engine. 

Allis  Pumps. 

Atlantic  City  Station. 

Beam  Engines. 

Beam  Engine  of  1881. 

Belmont  Worthington  Pump. 

Berlin  Water  Works. 

Birmingham  Water  Works. 

Birmingham,  Ala.,  Worthington  Pump. 

Boston-Leavitt  Pump. 

Boston  Allis  Pump. 

Brooklyn  Engine. 

Brooklyn  Pumps  of  1870. 

Brooklyn  Pumps. 

Brooklyn,  Davison  Pump. 
Brooklyn  Water  Works. 
Buda  Pesth  Water  Works. 
Buffalo  Holly  Pump. 
Cambridge,  Mass.,  Water  Pipe. 
Cast  Iron  Pipe. 
Chicago  Pumps  of  1866. 


Engng.,  41:  33. 
N.E.W.W.,  13:  172. 
Eng.  Rec.,  48:  215. 
Engng.,  i:  149. 
Engng.,  31:  64. 
Engng.,  15:  411. 
Engng.,  10 :  260. 
Engng.,  40:  298. 
Eng.  News,  27:  368. 
Eng.  News,  28:  578. 
N.E.W.W.,  June,  1901. 
Engng.,  i:  250. 
Engng.,  9:  369. 
Engr.,  May  i,  1904;  Engng., 

9:  369. 

Eng.  News,  12:  220. 
Eng.  News,  25:  225. 
Engng.,  39:  528;  574,  623. 
Engng.,  28:  365. 
N.E.W.W.,  n:  121. 
N.E.W.W.,  n:  27. 
Engng.,  7:  242.. 


690 


BIBLIOGRAPHY 


Chicago  Water  Works  of  1875. 
Chicago  Pumping  Station. 

Chicago  Allis  Pumps. 

Cincinnati  Pump. 

Cleveland  Plant. 

Cornish  Pump  of  1881. 

Columbus  Gaskill  Pump. 

Compound  vs.  Triple  Expansion 

Compensated  Gear. 

D'Auria  Pump. 

Davy  Differential  Pump. 

Dean  High  Duty. 

Development  and  Peculiarities  of  Water  Works 

Pumps,  by  I.  H.  Reynolds. 
Deep  Well  Plant. 
East  Jersey  Water  Supply. 
English  Beam  Fly  Wheel  Engine. 
Fall  River  Davison  Pump. 
Foreign  Water  Supply. 
Garden  City  Station,  Air  Lift. 
Gas  Engine  Plant. 
Hannibal  Allis  Pump. 
Horizontal  Compound  Pump. 
Kley,  Flywheel  Beam  Pump. 
Lambeth  Beam  Engine. 
Lardner's  Point  Station. 
Lawrence  Leavitt  Pump. 
Leavitt  Pumps. 
Memphis  Vertical  Worthington  Pump. 

Milwaukee  Allis  Engine. 
Moreland  Pump. 

Moreland  and  Thompson  Pump. 

Mystic  Pump.  -  , 

Newark  Worthington  Pump. 

New  Bedford  Pumps. 

New  London  Pump. 

Ottumwa,  Iowa. 

Paris  Farcot  Pump. 

Pawtucket  Pump. 

Peoria  Plant. 

Philadelphia  (see  Belmont,  Lardner's  Point). 

Philadelphia  Water  Works  of  1868. 

Port  Washington,  N.  Y. 

Prague  Water  Works. 

Present  Pumping  Engine  Practice,  Reynolds. 


Engng.,  19:  30. 

Eng.  News,  51:  485;  Eng 

Rec.,  48:  120. 
Eng.  News,  23:  506. 
Engng.,  3:  53*- 
Eng.  Rec.,  49:  348. 

Engng-,  32:  2O9- 
Eng.  News,  n:  118. 
N.E.W.W.,  13:  218. 
Engng.,  44:  349,  409. 
A.W.W.A,  1905,  304. 
Engng.,  22:  421. 
Eng.  News,  29:  137. 

A.S.C.E.,  1904,  No.  95. 
Engng.,  Feb.  25,  1910. 
N.E.W.W.,  8:  18. 
Engng.,  44:  457. 
Eng.  News,  10:  423. 
N.E.W.W.,  9:  109. 
Eng.  Rec.,  59:  535. 
Eng.  Rec.,  53:  196. 
Eng.  News,  14:  142. 
Engng.,  24:  10. 
Engng.,  15:  50. 
Engng.,  2:  273. 
Eng.  Rec.,  52:  315. 
Engng.,  27:  58. 
N.E.W.W.,  9:  163. 
Eng.  News,  22:    151;    26: 

233- 

A.E.S.,  14:  24: 
Engng.,  38:   319,  384,  472, 

520. 

Engng.,  6:  544. 
Eng.  News,  32:  176. 
Engng.,  10:  224. 
Eng.  Rec.,  49:  797. 
N.E.W.W.,  7:  148. 
Eng.  Rec.,  53:  430. 
Engng.,  26:  70. 
Engng.,  28:  189. 
Eng.  News,  28:  58. 

Engng.,  6:  267. 
Eng.  Rec.,  52:  205. 
Engng.,  44:   200,  396. 
N.E.W.W.,  13:  172. 


BIBLIOGRAPHY 


691 


Pumping  Machinery  for  Water  Works,  by  F.  H. 

Pond. 

Pipes,  see  Syracuse,  Cast  Iron,  East  Jersey,  Cam- 
bridge. 

Richardson  Triple  Expansion  Pump. 
Sidney  Pump. 
St.  Louis  Beam  Engine. 
St.  Louis  Pumps  of  1874. 
St.  Louis  Water  Works  of  1894. 
St.  Louis  Plants. 
St.  Paul  Allis  Engine. 
Small  Pump  Plants,  Farr. 
Syracuse  Steel  Pipe. 
Schenectady  Plant. 
Toledo  Water  Works. 

Vienna  Exposition  Pump. 

Vienna  Water  Works  Pump. 

Vertical  Pump. 

Washington,  D.  C,  Plant. 

Water  Supply  for  Small  Cities. 

Water  Works,  Ancient  and  Modem. 

Water  Works  Pumps  of  1875,  Sizes,  Costs,  Duty. 

Windsor,  Ont,  Plant. 

Worthington  High  Duty  Pump  in  England. 

Worthington  Pump. 

Zurich  Plant. 


Eng.  News,  13:  340. 


Engng.,  50:  158. 
Engng.,  24:  10;  42:  574. 
Eng.  News,  16:  93. 

Engng.,  31 :  143,  200. 
A.E.S.,  1894. 
Eng.  Rec.,  49:  700. 
Eng.  News,  15:  13. 
A.W.W.A.,  1907. 
N.E.W.W.,  8:  40. 
Eng.  Rec.,  51:  640. 
Eng.  Rec.,  52:  377;  Power, 

Feb.  21,  1911. 
Engng.,  16:  242. 
Engng.,  28:  432. 
Engng.,  39:  485. 
Eng.  Rec.,  53:  64. 
N.E.W.W.,  5:  83. 
Engng.,  21 :  502;  22:  7 
A.S.C.E,  4:  369. 
Engr.,  Nov.  16,  1903. 
Eng.  News,  21:  87. 
N.E.W.W.,  13:  229. 
Eng.  News,  32:  34. 


INDEX 


Absolute  path  of  water,  628 

Acceleration,  190 

Acceleration    and    velocity    curves, 

674 

Acceleration  diagram,  241 
Acceleration  of  parts,  333 
Admiralty  pumps,  137 
Ahrens  pump,  277 
Air  chambers,  242 

advantage  of,  243 

pressure,  247 

size  of,  248 

suction,  305 
Air  compressor,  520 
Air-lift  pump,  122 
Air-lift       pumps     and      pneumatic 

pumps,  512 
Air-lift  well  tops,  514 
Air  valve,  489 
Air  pump,  36 

Alberger,  465 

cards,  459 

dry,  464 

Edwards,  461,  462 

Mullen,  463 

navy,  460 

Air  pumps  for  beer  racking,   498 
Air  pumps,  size  of,  458 
Aix-la-Chapelle  mine  pump,  659 
Alberger  Pump  Co.,  575,  650 
Alberger  centrifugal  condenser,  583 
Alberger  multistage  turbine  pump, 

576 
Alberger    standard     volute    pump, 

vertical  shaft,  581 
Alberger    two-stage    volute    pump, 

58i 
Allis-Chalmers,  112 


Allis-Chalmers    duplex  Riedler  ex- 
press pump,  163 
Allis-Chalmers  pump,  438,  590 
Allis-Chalmers  Riedler  valve,   294 
Allis  pumps   at  the  Baden  station, 

_  435 

Allis  pumps  at  Bissel's  Point,  436 
Allis  pump  of  Milwaukee,  286 
Allis  screw  pump,  116 
American  Fire  Engine  Go's,  metro- 
politan engine,  277 
Andrews,  45,  46 
Archimedean  screw,  14 
Area  curve  through  bucket,  627 
Areas    at    entrance    and    discharge, 

608 

Area  scales,  339 
Arrangement  of  pump,  318 
Arrangement  of  vanes,  608 

Bach,  208,  210 
Back  vanes,  628 
Balancing  pumps,  400 
Ballast  pump,  141 
Ball  valve,  290 
Barlow's  formula,  309 
Barr,  3*6 

Bearing  power  of  soil,  402 
Bearings,  391 

design  of,  389 
Behrens  pump,  113 
Beighton,  Henry,  19 
Belmont  water- works,  77 
Belted  volute  pump,  567 
Benjamin,  C.  H.,  562 
Bibliography,  679 
Bilge  pump,  141 
Blake,  45,  62 

693 


694 


INDEX 


Blake  pump,  174,  176,  177 

Boiler-feed  pump,  134 

Borsig  valve,  301 

Boulton,  Mathew,  36,  39 

Box  end  design,  386 

Braithwaite,  52 

Brass  liner,  266 

Brooklyn     high     pressure     station, 

596,  655 

Brooklyn  station,  440 
Brotherhood  engine,  119 
Bucket,  4,  260 
Bucket,  Roman,  10 
Bucket  pump,  129,  130 
Buffalo   balanced  two-stage  pump, 

587 

Buffalo  pump,  587,  588 
Buffalo    vertical     underwriter     fire 

pump,  589 
Bull,  Wm.,  40 
Bull  Cornish  engine,  67 
Burnham  pump,  183,  184,  281 
Burnham  pump  valve,  300 
Bushing  rings,  565 
Butterfly  valves,  204,  289 

Cage,  pressure  valve,  300 

Calibrated  nozzle,  590 

Galley,  John,  30 

Cameron  pump,  165,  166,  169 

Cameron  valve,  292 

Cards,  combined,  334 

Carpenter,  403,  411 

Central  outside  packing,  132 

Central  Park  Avenue  station,  434 

Centrifugal  pressure  head,  539 

Centrifugal  pumps,  535 

Centrifugal  pump  dimensions,  570 

Centrifugal     pump,     efficiency    cf, 

594 

Centrifugal  pumps  for  special  pur- 
poses, 646 

Chaplets,  10 

Check  valve,  490 

Chesney,  Col.,  6 

Cincinnati  pump,  286,  428 

Circular  arc  vanes,  618 

Clack  valves,  208,  220 


Clack  valve,  hinged,  288 

leather,  288,  289 

double,  289 

metal,  289 

Clack  valves,  rectangular,  281 
Clearance,  235,  257 
Cloth  dryer,  38 
Cock  valves,  204 
Combined  cards,  334 
Combined    entrance    and    discharge 

diagram,  549 

Compound  direct-acting  pump  dia- 
gram, 676 
Compound     outside     end     packed 

pressure  pump,  144 
Compound  pumps,  71 
Compound    outside    center   packed 

boiler  feed  pump,  135 
Compressor  with  clearance,  522 
Concrete  foundations,  402 
Condensers,  36,  406 

Buckley,  86 

Connections,  water  pipe,  319 
Conical  valve,  289,  290 
Connecting  rods,  361,  367 

design,  370 

marine  end,  369 

strap  end,  368 
Connecting-rod  box,  382 
Connecting-rod  ends,  design,  386 
Consolidated  Cal.  shaft,  66 1 
Controllable  valve  arrangement,  489 
Cooling  water,  quantity,  457 
Core  hole  cap,  310 
Corliss,  74,  78,  79,  81,  95 
Corliss  valves,  95,  96 
Cornish,  40,  41,  42,  43 
Cornish  engine,  86,  87 
Cornish  valves,  95 
Coupling,  590 

Covers,  manhole  or  handhole,  310 
Crank  pin  design,  374 
Critical  speed,  636 
Cross-head  design,  370 
Cross-head,  274,  361,  363,  364 
two  rod,  365 
pump  end,  365 
cross-head  pin  design    3  7 1 


INDEX 


695 


Cup  leathers,  265 

Cup-leather  packing,  261 

Ctesibius,  15 

Curves  of  area,  622 

Curves  of  quantity,  622 

Curves   of  variation   of  coefficient 

of  friction,  561 
Cylinder,      design     of,     307,     310, 

355 

Cylinder  openings,  310 
Cylinder  ratio,  329 
Cylinders,  sizes,  330 

Cylinders,  steam,  348 

triple  expansion,  346 

Dalby,  W.  E.,  400 

d'Auria,  Luigi,  104 

Davey,  Henry,  105 

Davey  compensator,  106 

Davies,  J.  D.,  98 

Davison,  97 

Davidson  pump,  180,  181 

Davidson    vertical    duplex    pump, 

138 
Davidson  horizontal  steam  cylinder 

for  deep-well  pump,  156 
Dean  Bros,  pump,  186,  187 
Deane  pump,  176,  177,  178 
Deane  triplex  vertical  single-acting 

power  pump,  161 
Deep- well  pump,  153,  278 
de  Lorme,  Philibert,  126 
Desagulier,  33 
Design  of  bearings,  389 
Design  of  centrifugal  pumps,  602 
Design  of  cross-head,  connecting- 
rod  shaft,  370 

Design  of  cylinder,  307,  355 
Design  of  parts,  260 
Design  of  piston  rod,  361 
Design  of  springs,  316 
Design  of  steam  piping,  403 
Details  steam  end,  344 
Diagrams  of  velocity,  237,  239,  241 

acceleration,  241 

space,  238,  239,  241 
Diagram,  scales,  340 
Diameters  of  impellers,  610 


Differential  bucket  pump,  131 

Differential  plunger  pump,  131 

Diffuser,  536,  574,  631 

Diffuser  vanes,  536 

Diffusion  chamber,  630 

Direct-acting  water  works  pumps, 
152 

Discharge  chamber,  265 

Discharge  cone,  488 

Discharge,  rate  of,  192 

Discharge,  velocity  of,  189 

Disc  valve,  291 

Doon,  ii 

Double-acting  plunger  pump,   131 

Double-beat  valve,  296 

Double-flow  pump,  537 

Double-flow  volute  pump,  568 

Double  leather  packing,  262 

Double-ported  valve,  294,  296 

Drain  cocks,  493 

Dredging  pumps,  652 

Driving  valves,  mechanism,  346 

Dry  air  pumps,  464 

Dry  dock  units,  647 

Dunkerley's  values  for  shafts,  643 

Duplex  pumps,  64,    133,    134,   344 

Duty  trial  of  test  of  pumping  en- 
gine, 410 

Duties,  43 

Dynamics  of  steam  end,  324 

Dynamics  of  water  end,  189 

Effect   of  changing   quantity   with 

fixed  speed,  555 
Efficiency,  418 

Efficiency  of  centrifugal  pump,  594 
Efforts,  tangential,  338 
Electric  sinking  pump,  150 
Emerson  pump,  508 
Engine,  Cornish,  86 

horizontal  fly-wheel,  87 

rotary  36,  38 

double-acting,  38 

trunk,  38 

Ericsson,  Capt.  John,  52,  53 
Euler,  43 
Eve,  J.,  48,  49 
Expansive  use  of  steam,  36,  38 


696 


INDEX 


Explosion  pump,  128 
Express  pump,  163 
Express  pumps  in  series,  663 

Fairbanks-Morse     deep-well    pump, 

I53.  J54 

Fairbanks-Morse  pump,    270,    271 
Fielding,  105 
Filling  ring,  565 
Fire  pump,  Ahrens,  277 

Amer.  Fire  Engine  Co.,  277 

double-acting,  275 

Metropolitan,  276 
Fire  pumps,  workmanship,  468 

duplex  only,  468 

sizes,  468 

capacity,  469 

speed,  469 

capacity  plate,  470 

strength  of  parts,  470 
Fire  engine,  Merryweather's,  299 
Fire  engine,  shop  inspection,  470 

steam  cylinders,  471 

steam  ports,  471 

steam  clearance,  471 
Fire-pump,  French,  51,  52 
Flanges,  311 
Flange,  shrunk,  404 

F,  screwed,  404 

F,  welded,  404 

Flash  wheels,  or  scoop  wheels,  55 
Fluctuation  factor,  341 
Fly-wheel  design,  393 
Fly-wheel  energy,  340 
Fly-wheel,  sectional,  394 

size,  342 

Follower  plate,  266 
Foot  bearing,  536 
Foot  or  suction  valve,  278 
Foot  valve,  305,  306,  492 
Forces    in    direct- acting    rod  mine 

pump,  670 
Force  pump,  15 

Forms  of  centrifugal  pumps,    565 
Forms  of  impeller,  613 
Foundations,  of   concrete,  of  piles, 

402 
Fountain,  16 


Fowler,  Geo.,  651 

Frame  design,  400 

Frames  for  vertical  engine,  400 

Frames,  391,  400 

Francis,  593 

Friction  losses,  197 

Friction  in  pipes,  200 

Friction  at  bearings  and  stuffing- 
boxes,  559 

Friction  of  water  on  back  of  im- 
peller, 563 

Frizell,  J.  B.,  122 

Gas  engine  for  deep-well  pump,  156 

Gaskell,  H.  F.,  82,  83 

Gate  valve,  204 

Gate  valve,  Ludlow,  405 

Gauge,  mercury  steam,  38 

water,  38 

Gelpcke-Kugel,  590 
General  service  pumps,  140 
Giffard,  H.  J.,  125,  126 
Globe  valve,  404 
Gould  steam  fire  pump,   119 
Governor,  engine,  38 
Graff,  Fred,  70,  71,  72 
Graphical   method    of    centrifugal 

pump  design,  611 
Green,  D.  M.,  Prof.,  96 
Guest,  J.  J.  381 
Gutermiith  valve,  301 
Gwynne,  John,  45 
Gwynne,  J.  and  H.,  no 

Hague,  C.  A.,  94,  95 
Valve  method,  219 
Hammer,  steam,  38 
Hand  fire  pump,  12 1 
Hand  holes,  270 
Harris,  E.  G.,  124 
Harris  pneumatic  pumps,  425,  531 
Hartmann  and  Knoke,  213,  215 
Head  limit,  233 
Heat  determination,  413 
Helical  pump,  118 
Helical  rings,  290 
Helicoidal  pump,  118 
Heisler  pump,  105 


INDEX 


697 


Hero,  51 

Hero  of  Alexandria,  14 

Hesse,  F.  G.,  563 

High-duty  pumps,  428 

High-pressure  pumping  stations: 

Brooklyn,  655 

New  York  city,  656,  657 

Philadelphia,  654 
Hinged  clack  valve,  288 
Holly,  Birdsill,  68 
Holly  Manufacturing  Co.,  82 
Holly  pump  at  Boston,  437 
Holly  pump  at  Washington,  D.  C., 

439 

Holly  triple  expansion  pump,  108 
Hornblower,  Jonathan,  39,  40 
Horizontal  fly-wheel  pumping  engine, 

87 

Horse-power,  167 
Hose  valves,  488 
Humphrey,  explosion  pump,  128 
Humphrey,  H.  A.,  126 
Hydraulic  ram,  53,  500 
Hydraulic  presses,  310 
Hydraulic  pressure  pumps,  496 

Impeller,  536,  565 
Indicator,  steam-engine,  38 
Indicator  cards,  actual,  331 

combined,  335 
Inertia,  effect  of,  230 
Interference  of  blunt  vanes,  554 
Injectors,  125 

Injector  and  pulsometer,  503 
Involute  curves,  616,  617 

Jackets,  36,  418 
Jacobus,  J.  S.,  339 
Jeanesville  mine  pump,  668 
Jeanesville  valve  pots,  669 
Jordan,  Johann,  43 
Joseph's  well,  7 

Katweh,  10 

Kent,  411 

Kent's  pocket-book,   417 

Knsass,    Strickland    L.,    505,    506 


Knowles,  62 

Knowles  express  pump,   665,   666 
Knowles  pump,   172,   173,   174 
Knowles  simplex  pump,  134 
Knowles  steam  pump  works,  665 
Knowles  underwriters'  pump,  142, 

144 
Knowles    vertical    duplex    electric 

sinking  pumps,   149 

Lagging,  36 

Lambeth  Water  Works,  88 

Lame  formula,  309 

Larner  pump,  584 

Larner,  C.  W.,  582 

Lat,  4 

Latha,  10 

Lawrence  pump,  283 

Leakage,  557 

Leather  clack  valve,  288,  289 

Leavitt,  E.  D.   jr.,  70,  71,    77,    78, 

107,  281,  347 
Lewecki,  564 
Liner,  independent,  351 
Lloyd,  45 

London   Bridge  Water  Works,    18 
Loss  in  bends,  202    . 
Loss  in  passages,  199 
Loss  in  pipes,  198 
Loss  in  valves,  202 
Loss  due  to  inertia,  206 
Loss  due  to  sudden  contraction,  205 
Loss  due  to  velocity  changes,  204 
Losses  in  centrifugal  pumps,  552 
Loss  through  valve,  198,  208 
Ludlow  gate  valve,  405 
Luitwieler  pump,  155,  157,  158 

MacFarland's  rotary  pump,  114 
McCarty,  43,  44,  45 
Maltby,  F.  B.,  651 
Manhole    or    handhole   cover,  309 
Manhole  cover,  288 
Mair,  J.  I.,  101 
Marine  boiler  feed  pump,  136 
Marine  end,  369 

Marsh  pump,   265,    266,    267,   268, 
269 


698 


INDEX 


Marsh  steam  pump,  169,  170,  187 

Marsh  valve,  293 

Massachusetts  pump,  44,  45,   no 

Mean  pressure,  326 

Measuring  pressure,  412 

Measuring  water  pumped,  411 

Mechanical  efficiency,  325 

Mental,  10 

Merriman's  hydraulics,    593 

Merryweather's  fire  engine,  299 

Metal  clack  valve,  289 

Metal  disc  valve,  292 

Metallic  packing,  355 

Method  for  volute  pumps,  613 

Mexican   union   mine   pump,    659 

Milk  pump,  266 

Milwaukee  pump,  285 

Mine  pumps,  146,  659 

Aix-la-Chapelle,  659 

consolidated  Cal.  shaft,  66 1 

express  pumps,  663 

German,  66 1 

Jeanesville,  668 

Mexican  union,  659 
Mine  pump,  Scran  ton  pattern,  149 
Mixed  flow  pumps,  619 
Modern  fire  engine,  American,  La 

France  Co.,  122 
Modern  forms  of  pumps,  129 
Montgolfier,  54,  55 
Moody,  Prof.  L.  P.,  599 
Moreland's  compound  steam  end, 

80,  81 

Moreland,  Rich.,  68 
Moreland,  96,  in 

Sir  Saml.,  24 

Morrys,  Peter,  18,  20,  21  - 
Mot,  n 

Moving  parts,  weights  of,  332 
Multistage  compression,  522 
Multiple  expansion,  326 
Muntz  metal,  314 

Neumann,  547 
Neumann's  curves,  548 
Newcomen,  Thos.,  28,  29,  30,  33 
New   York   high-pressure    station, 
656,  657 


Non-aligning   ring   oiling  bearing, 

559 

Norberg   quadruple-expansion    en- 
gine, 107 

Noria,  2,  4 

Number  of  revolutions,  255 

Number  of  stages,  602 

Odell,  564 

Oil  cellar,  566 

Oil  rings,  566 

Oil  thrower,  566 

Ontario  pump,  284 

Openings,  310 

Outboard  bearing,  566 

Outside  packed  plunger  boiler  feed 

pump,  136 
Outside  packed  pump,  169 

Packings,  36,  314 
Packing,  U-leather,  263 

plunger,  260,  263 

cup  leather,  261 

double  leather,    262 
Paecottah,  4 
Pallets,  10 

Papin,  Denys,  27,  43 
Pappenheim,  47 
Parabolic  vanes,  615 
Paris  pump,  88 
Parts,  design  of  260, 
Peabody,  409 
Persian  wheel,  7 
Philadelphia  high-pressure  station, 

654 

Phillips  rotary  pump,  115 
Piles,  402 
Pipe  friction,  200 
Pipe  loss,  198 
Pipes,  size  of,  256 
Piston,  260,  353,  354 

pressure,  197 
Piston  proportions,  358 
Piston  pump,  129,  130 
Piston  ring  design,  358 
Piston  rods,  314 
Piston  rod  design,  361 
Pistons,  water,  313 


INDEX 


699 


Piston  with  sectional  ring,  354 
Pitot  tube,  412,  590 
Plunger,  packed,  260 
Plunger  and  ring,  260,  271 
Plunger  and  ring  packing,  132 
Plungers,  outside  packed,  274 

center,  oustide  packed  274 
Plunger  packing,  260,  263 
Plunger  pump,  129,  130 
Pohle,  Julius,  122 
Pointed  vanes,  554 
Points  of  leakage,  558 
Power  pumps,  horizontal  duplex, 

1 60 

Power  head  deep- well  pump,   162 
Preparation  for  test,  419 
Press,  letter  copy,  38 
Prescott    duplex     outside  -  packed 
plunger    pot-form    boiler-feed 
pumps,  138 

Prescott  steam  sinking  pump,  150 
Pressure  cylinder,  310 
Presses,  hydraulic,  316 
Press  and  pump,  500- ton,  497 
Pressure     on     discharge     without 
air  chamber,  237 

with  air  chamber,  243,  247 
Pressure,  measuring,  412 
Pressure    pumps,      145,      279,     280, 

281 

Pressure,  resultant,  238,  336 
Pressures,  terminal,  328 
Pressure  valves,  298,  299 
Priming,  489 
Priming  tank,  492 
Pulsometer,  126 
Pump,  air-chamber,  303,  304 
Pump  arrangement,  318 
Pump,  dredging,  652 
Pumping  engines,  test  of,  410 
Pump,  Fairbanks-Morse,  270,   271 

Worthington,  272,  273 
.  Burnham,  281 

Marsh,  265 

milk,  266 
Pumps,  condenser,  456 

combined  air  circulating,  457 
Pumps,    pressure,    279,    280,    281 


Pump  with  air  chamber,  242 
Pump  rods,  278 
Pump,  sewage,  282 

Cincinnati,  286 

Snow,  287 

Lawrence,  282 

Ontario,  284 

Milwaukee,  285 

Riedler,  451,  452 
Pump  specifications,  underwriters', 

468 
Pump,  sewage,  466 

underwriters',  467 
Pumps,  turbine,  649 
Pumping  machinery,  special,  456 

Quadruplex  pump,  69 
Quimby  screw  pump,  127 

Railroad  pump,  277 

Ramelli,  Agostino,    18,  20,  23,  28, 

47 

Ramseye,  David,  18,   23 
Rankine,  381 
Rateau,  584 
pump,  585 

Reciprocating  pump,  51 
Reciprocating    parts,     weight    of, 

333 

Rectangular  clack  valves,  281 
Reheaters,  329 
Relief  valve,  488 
Resultant  pressure,  235,  336 
Return  chamber,  588 
Revillion,  51,  52    . 
Revolutions,  number  of,  258 
Reynolds  pump,  92,  95 
Reynolds,  I.  H.,  valve  method,  219 
Riedler  pump,  163,  164,  451 
Riedler  valve,  294 
Results  of  actual  test,  419- 
Roebuck,  Dr.,  34 
Root  rotary  pump,  116 
Rotary  pumps 

Behrens,  113 

MacFarland's,  114 

Phillips;  114 

Root,  116 


700 


INDEX 


Rotary  pumps 

Silsby,  115 

"VYilkins,  115 

Rotary  pump,  Ramelli's  sixteenth 
century,  47 

Trotter's,  four-bladed,  50 

Watt's  Eve's,  48 
Rubber  check  valve,  491 
Rubber  valve,  293 
Runner,  536 

Safety  valve,  488 

Sakias,  7 

Savannah  pump,  62,  63 

Savery,  Thos.,   25,   26,   27,   28,  33 

Scales  of  diagrams,  340 

Scales,  area,  339 

Schwartzkopff,  585 

Louis,  in 
Scoop  wheels,  55 
Screw,  Archimedean,  14 
Sectional  fly-wheel,  394 
Section  of  impeller,  614 
Section    of   mixed    flow    impeller, 

620  - 
Self-aligning    ring    oiling    bearing, 

560 

Sellers  injector,  503 
Serviere,  46,  49,  116 
Sewage  pump,  282,  466 
Shaft  design,  370,  633 
Shaft  sketch,  376 
Shields,  Geo.,  67 
Shrunk  flange,  404 
Shadoof,  2 

Silsby  rotary  pump,  115 
Simplex  boiler  feed  pump,  135 
Simplex  pumps,  133,   165,   172 
Simpson's  pumping  engine,  88 
Simpson  &  Co.,  649 
S;nking  pump,  148 
Size  cylinders,  330 
Size  of  air  chamber,  248 
Size  of  air  pumps,  458 
Size  of  pipes,  254 
Size  of  valves,  213 
Smeaton,  John,  31,  33 
Snow  compound  pump,  108 


Snow  pump  end,  287 
Soil,  bearing  power,  402 
Somerset,  Edward,  23,  24 
Space  diagram,  238,  239,  241 
Specific  speed,  598 
Specific  speed  curves,  603 
Speed,  critical,  636 
Speed,  specific,  598 
Spindle,  valve,  292 
Split-casing  pumps,  572 
Spring-controller  valves,  270 
Spring  design,  316 
Springs,  316 
Stations : 

Brooklyn,  440 

Cincinnati,  428 

Ferrara,  648 

High  Bridge,  444 

Jersey  City,  454 

Kinnickinnic  River,  445 

Lardner's  Point,  447 

Memphis,  441 

Zurich,  443 
Steam  cylinder,  348 
Steam  end  details,  344 
Steam  end  dynamics,  324 
Steam  piping,  403 
Steam  piston,  352 
Steam  valve,  345 

Steam  valve  of  duplex  pump,  344 
Stodola,  564 
Strainer,  305,  306 
Strap  end  connecting  rod,  368 
Strap  end  design,  388 
Stroke  ratio,  255 
Stuffing-box,  262,  263,  266,  315 
Suction  air  chamber,  305 
Suction  bearing,  566 
Suction  chamber,  265 
Suction  head,  565 
Suction  pipe,  536 
Sulzer,  583 
Sulzer  pump,  584 
Swape,  4 

Table  of  lost  head,  607 

Taboot,  7 

Tangential  effort  construction,  337 


INDEX 


701 


Tangential  efforts,  338 

Tank  pumps,  140 

Taper  threads,  278 

Terminal  pressures,  328 

Test  at  Lardner's  Point,  419 

Test  curves,    595,.  596,     597,    607, 

667 

Test  data,  417 
Test  for: 

acceptance,  493 
internal  friction,  494 
strength  and  tightness,  494 
internal  leakage  or  slip,  495 
maximum  working  pressure,  495 
maximum  delivery,  495 
Testing  centrigufal  pumps,  590 
Test  of  air  lift  pump,  515,  518 
Test  of  internal  friction,  494 
Test  of  pumping  engines,  411 
Test    of     strength    and    tightness, 

494 

Test  precautions,  416 
Test  results,  453 
Thames-Ditton,  71 
Thompson,  David,  68 
Thurston,  R.  H.,  107 
Thrust  bearing,  566,  575 
Tobin  bronze,  143,  314 
Trevithick,  Rich.,  40 
Triangular  weir,  594 
Triple  expansion  pump,  95,  96 
Triple  expansion  cylinders,  346 
Triplex  pump,  133,159 
Trirotor  volute  pump,  572 
Trombone  frames,  273 
Trotter,  John,  49 
Turbine  pump,  573 
Two-stage  pump,  112 
Twin  cylinder  casting,  310 
Tympanum,  13 

Underwriters'     fire    turbine    type, 

579 

U-leather  packing,  263,  265 
U-leathers,  265 
Underwriters'  pump  specifications, 

468 
Underwriters'  pumps,  141 


Unwin,  358,  396,  564 

Valve  acceleration,  216 
Valve,  air,  489 
Valve  backing,  292 
Valve,  ball,  290,  291 

Borsig,  301 
Valve  box,  286,  297 
Valve,  Burnham,  299,  300 

butterfly,  204,  289 
Valve  cage,  286 
Valve,  Cameron,  292 
Valves,  check,  490. 

clack,  208,  220 

cock,  204 
Valve,  conical,  289 

dash  relief,  74 
Valve  deck  c.over,  265 
Valve  deck-plate,  265 
Valve  design,  Hague's  method,  219 

Reynolds'  method,  219 
Valve  dimensions,  220 
Valve,  disc,  291 
Valve  discs,  317 
Valves,  double  beat,  296 
Valve,  double-ported,  294,  296 

foot,  305,  492 

foot  or  suction,  378 

gate,  204 
Valve  gear,  349 

pressure  control,  350 
Valve,  globe,  404 

Gutermuth,  301 
Valve  handle,  490 
Valves,  head,  351 
Valve  loss,  202 
Valve,  Ludlow  gate,  405 

Marsh,  293 
Valve  movement,  217 
Valve,  multiported,  224 

positive  water  and  steam,  59 
Valve  pot,  300 
Valve,  pressure,  .298 
Valves,  rectangular  clack,  281 
Valve,  resistance,  229 
Valves,  relief,  488 
Valve,  relief  motion,  60 
Valves,  resistance  of,  224,.  226 


INDEX 


Valve,  Riedler,  294 
Valve  rod  yoke,  345 
Valve,  rotary  steam,  74 

rubber,  293 

safety,  488 

Valves  and  springs,  316 
Valves,  hose,  488 
Valve  spindle,  292 
Valves,   spring-controlled,    270 
Valve,   spring- thrown,    57,    59 
Valves,  steam,  345 
Valve  weighted,  295 

Witting's,  298 
Vane  curves,  613 
Velocity  diagram,  238,  239,  241 
Velocities  of  discharge,  199,  541 
Velocities  at  entrance,   541 
Venturi  meter,  412,  590 
Vertical  pump,  101 
da  Vinci,  Leonardo,  17,  28 
Vitrio,  126 

Volute  casing^ 536,  565,  632 
Volute  centrifugal  pump,  565 

Warren  steam  pump,  172,  173 
Warren  Steam  Pump  Co.,  499 
Waterbury  Farrel  Foundry  &  Ma- 
chine Co.,  497 
Water  cyliners,  260 
Water  end,  dynamics,  189 
Water-pipe  connections,  319 
Water  pistons,  313 
Water  works  pump,  151,  153 
Water  works,  428 

Allegheny,  93 

Berlin,  89 

Brockton,  108 

Brooklyn,  63 

Buffalo,  70 

Cambridge,  63 

Charleston,  63,  67 

Cincinnati,  67,^428 

Dunkirk,  68 

Eastbourne,  68,  8 1 

East  London,  89 

Fall  River,  101 

Hannibal,  96 

Lambeth,  89 


Water  works,  Lardner's  Point,  445 

Lawrence,  78 

Lockport,  68,  94 

Lynn,  70,  71 

Milwaukee,  93 

Newark,  77 

Pawtucket,  78,  79 

Philadelphia,  67,  72,  77 

Pittsburg,  107 

Providence,  88 

Rochester,  70 

St.  Maurs,  88 

Saratoga  Springs,  82 

Savannah,  63 
Watt.J.,  21,33,  38 
Weighted  valve,  295 
Weights  of  moving  parts,  332 
Weight     of      reciprocating     parts, 

333 

Weise  &  Monski  pump,  585 
Welded  flange,  404 
Wheeler  Condenser  &  Engineering 

Co.,  456 

Wheeler  system,  517 
Whirlpool  chamber,  537 
Whitehurst,  53 
Wilkin,  J.  T.,  115 
Witting's  metallic  valves,  298 
Wood    45-inch    centrifugal   pump, 

650 

Wood  propeller  pump,  117 
Wood,  R.  D.,  649 
Worcester,  Marquis,  23,  24 
Work  in  centrifugal  pumps,  543 
Worthington  ballast  pump,   143 
Worthington    beer-racking    pump, 

498 
Worthington    eight-stage    turbine 

mine  pump,  577 
Worthington       fireboat       turbine 

pump,  578 
Worthington  four-stage  boiler  feed 

pump,  579 
Worthington    high    speed     pump, 

580 

Worthington  mine  pump,   147 
Worthington    packed-plunger 

pump,  139,  148 


INDEX 


703 


Worthington  piston  pump,  141 

Worthington,   58,   59, 

pump,  57,  60,  61,  62,  63,  67, 
73,  75,  76,  77,  85,95,  96,  99, 
101,  137,  139,  141,  165,  272, 

273 
Worthington   pumps,    Fall   River, 

Mass.,  434 
Worthington  ten-stage  pump,  577 


V/orthington    three-stage    turbine 

pump,  586 
Worthington  trirotor  volute  pump, 

572 
Worthington  turbine  sinking  pump, 

580 
Worthington  volute  pump,  562 

Zig-zag  balance,  1 1 


This 


DEC  121 
MAY  241953 


Jfi 


«fl 


Enci 


neering 


UNIVERSITY  OF  CALIFORNIA  UBRARY    *Jl:  -,•' 


